Industrial machinery repair: best maintenance practices pocket guide

Industrial Machinery Repair: Best Maintenance Practices Pocket Guide Industrial Machinery Repair: Best Maintenance Pr...

Author: Ricky Smith | R. Keith Mobley President and CEO of Integrated Systems Inc.

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Industrial Machinery Repair: Best Maintenance Practices Pocket Guide


Ricky Smith and R. Keith Mobley

Amsterdam Boston London New York Oxford Paris San Diego San Francisco Singapore Sydney Tokyo

Butterworth–Heinemann is an imprint of Elsevier Science. Copyright © 2003, Elsevier Science (USA). All rights reserved. No part of this publication may be reproduced, stored in a retrieval system, or transmitted in any form or by any means, electronic, mechanical, photocopying, recording, or otherwise, without the prior written permission of the publisher. Recognizing the importance of preserving what has been written, Elsevier-Science prints its books on acid-free paper whenever possible. Library of Congress Cataloging-in-Publication Data Smith, Ricky. Industrial machinery repair : best maintenance practices pocket guide / Ricky Smith and Keith Mobley. p. cm. ISBN 0-7506-7621-3 (pbk : alk. paper) 1. Machinery–Maintenance and repair. 2. Industrial equipment–Maintenance and repair. I. Mobley, Keith. II. Title. TJ153.S6355 2003 621.8 16–dc21 2003040435 British Library Cataloguing-in-Publication Data A catalogue record for this book is available from the British Library. The publisher offers special discounts on bulk orders of this book. For information, please contact: Elsevier Science Manager of Special Sales 200 Wheeler Road, 6th Floor Burlington, MA 01803 Tel: 781-313-4700 Fax: 781-313-4882 For information on all Butterworth–Heinemann publications available, contact our World Wide Web home page at: http://www.bh.com 10 9 8 7 6 5 4 3 2 1 Printed in the United States of America

Contents Acknowledgments Chapter 1 Chapter 2

Introduction: Why Use Best Maintenance Repair Practices?

vii 1

Fundamental Requirements of Effective Preventive/Predictive Maintenance

10

Chapter 3

Maintenance Skills Assessment

26

Chapter 4

Safety First, Safety Always

50

Chapter 5

Rotor Balancing

57

Chapter 6

Bearings

71

Chapter 7

Chain Drives

120

Chapter 8

Compressors

133

Chapter 9

Control Valves

180

Chapter 10

Conveyors

203

Chapter 11

Couplings

215

Chapter 12

Dust Collectors

245

Chapter 13

Fans, Blowers, and Fluidizers

261

Chapter 14

Gears and Gearboxes

283

Chapter 15

Hydraulics

314

Chapter 16

Lubrication

327

Chapter 17

Machinery Installation

348

Chapter 18

Mixers and Agitators

353

Chapter 19

Packing and Seals

361

vi Contents

Chapter 20

Precision Measurement

386

Chapter 21

Pumps

395

Chapter 22

Steam Traps

432

Chapter 23

V-Belt Drives

441

Chapter 24

Maintenance Welding

460

Appendix A

539

Index

541

Acknowledgments Ricky Smith wants to offer his thanks to the following individuals who contributed to the writing of this book. Bruce Hawkins, Life Cycle Engineering; Darryl Meyers, former U.S. Army Warrant Officer; Steve Lindborg, Chemical Lime Company; Robby Smith (his brother), International Paper Corporation; and J.E. Hinkel, Lincoln Electric Company. Ricky also wants to thank Life Cycle Engineering, where he is currently employed, for the opportunity to write this book; Alumax–Mt. Holly—currently Alcoa– Mt. Holly—where he worked as a maintenance technician, for all the training and the chance to expand his knowledge; and Dr. John Williams, who always believed in him.



Ricky Smith and R. Keith Mobley

Amsterdam Boston London New York Oxford Paris San Diego San Francisco Singapore Sydney Tokyo

Butterworth–Heinemann is an imprint of Elsevier Science. Copyright © 2003, Elsevier Science (USA). All rights reserved. No part of this publication may be reproduced, stored in a retrieval system, or transmitted in any form or by any means, electronic, mechanical, photocopying, recording, or otherwise, without the prior written permission of the publisher. Recognizing the importance of preserving what has been written, Elsevier-Science prints its books on acid-free paper whenever possible. Library of Congress Cataloging-in-Publication Data Smith, Ricky. Industrial machinery repair : best maintenance practices pocket guide / Ricky Smith and Keith Mobley. p. cm. ISBN 0-7506-7621-3 (pbk : alk. paper) 1. Machinery–Maintenance and repair. 2. Industrial equipment–Maintenance and repair. I. Mobley, Keith. II. Title. TJ153.S6355 2003 621.8 16–dc21 2003040435 British Library Cataloguing-in-Publication Data A catalogue record for this book is available from the British Library. The publisher offers special discounts on bulk orders of this book. For information, please contact: Elsevier Science Manager of Special Sales 200 Wheeler Road, 6th Floor Burlington, MA 01803 Tel: 781-313-4700 Fax: 781-313-4882 For information on all Butterworth–Heinemann publications available, contact our World Wide Web home page at: http://www.bh.com 10 9 8 7 6 5 4 3 2 1 Printed in the United States of America

Contents Acknowledgments Chapter 1 Chapter 2


vii 1


10

Chapter 3


26

Chapter 4


50

Chapter 5

Rotor Balancing

57

Chapter 6

Bearings

71

Chapter 7

Chain Drives

120

Chapter 8

Compressors

133

Chapter 9

Control Valves

180

Chapter 10

Conveyors

203

Chapter 11

Couplings

215

Chapter 12

Dust Collectors

245

Chapter 13

Fans, Blowers, and Fluidizers

261

Chapter 14

Gears and Gearboxes

283

Chapter 15

Hydraulics

314

Chapter 16

Lubrication

327

Chapter 17

Machinery Installation

348

Chapter 18

Mixers and Agitators

353

Chapter 19

Packing and Seals

361

vi Contents

Chapter 20

Precision Measurement

386

Chapter 21

Pumps

395

Chapter 22

Steam Traps

432

Chapter 23

V-Belt Drives

441

Chapter 24

Maintenance Welding

460

Appendix A

539

Index

541

Acknowledgments Ricky Smith wants to offer his thanks to the following individuals who contributed to the writing of this book. Bruce Hawkins, Life Cycle Engineering; Darryl Meyers, former U.S. Army Warrant Officer; Steve Lindborg, Chemical Lime Company; Robby Smith (his brother), International Paper Corporation; and J.E. Hinkel, Lincoln Electric Company. Ricky also wants to thank Life Cycle Engineering, where he is currently employed, for the opportunity to write this book; Alumax–Mt. Holly—currently Alcoa– Mt. Holly—where he worked as a maintenance technician, for all the training and the chance to expand his knowledge; and Dr. John Williams, who always believed in him.

1


“Only Permanent Repairs Made Here” This book addresses, in a simplistic manner, the proper principles and techniques in “Best Maintenance Practices—Mechanical.” If these principles and techniques are followed, they will result in a serious reduction in “self-induced failures.” This book is a tool that should be carried and referenced by all mechanical maintenance personnel. A number of surveys conducted in industries throughout the United States have found that 70% of equipment failures are self-induced. Maintenance personnel who are not following what are termed “Best Maintenance Repair Practices” substantially affect these failures. Between 30% and 50% of the self-induced failures are the result of maintenance personnel not knowing the basics of maintenance. Maintenance personnel who, although skilled, choose not to follow best maintenance repair practices potentially cause another 20% to 30% of those failures. The existence of this problem has been further validated through the skills assessment process performed in companies throughout the state of Georgia. This program evaluated the knowledge of basic maintenance fundamentals through a combination of written, identification, and performance assessments of thousands of maintenance personnel from a wide variety of industries. The results indicated that over 90% lacked complete basic fundamentals of mechanical maintenance. This book focuses on the “Best Maintenance Repair Practices” necessary for maintenance personnel to keep equipment operating at peak reliability and companies functioning more profitably through reduced maintenance costs and increased productivity and capacity. The potential cost savings can often be beyond the understanding or comprehension of management. Many managers are in a denial state regarding maintenance. The result is that they do not believe that repair practices directly impact an organization’s bottom line or profitability.

2 Introduction: Why Use Best Maintenance Repair Practices?

More enlightened companies have demonstrated that, by reducing selfinduced failures, they can increase production capacity by as much as 20%. Other managers accept lower reliability standards from maintenance efforts because they either do not understand the problem or they choose to ignore this issue. A good manager must be willing to admit to a maintenance problem and actively pursue a solution. You may be asking, what are the “Best Maintenance Repair Practices”? Here are a few that maintenance personnel must know. (See Table 1.1.) Looking through this abbreviated Best Maintenance Repair Practices table, try to determine whether your company follows these guidelines. The results will very likely surprise you. You may find that the best practices have not been followed in your organization for a long time. In order to fix the problem you must understand that the culture of the organization is at the bottom of the situation. Everyone may claim to be a maintenance expert but the conditions within a plant generally cannot often validate that this is true. In order to change the organization’s basic beliefs, the reasons why an organization does not follow these best practices in the repair of their equipment must be identified.

“Only Permanent Repairs Made Here” A few of the most common reasons that a plant does not follow best maintenance repair practices are: 1

Maintenance is totally reactive and does not follow the definition of maintenance, which is to protect, preserve, and prevent from decline (reactive plant culture).

2

Maintenance personnel do not have the requisite skills.

3

The maintenance workforce lacks either the discipline or direction to follow best maintenance repair practices.

4

Management is not supportive, and/or does not understand the consequences of not following the best practices (real understanding must involve a knowledge of how much money is lost to the bottom line).

Table 1.1 Best maintenance repair practices

Maintenance task Standard

Required best practices

Lubricate Bearing

1 2 3

Lubrication interval: time based ± 10% variance.

4 Coupling Alignment

V-Belts

Align motor couplings utilizing dial indicator or laser alignment procedures. (Laser is preferred for speed and accuracy.) Straightedge method is unacceptable.

1

Measure the tension of V-belts through tension and deflection utilizing a belt tension gauge.

1

2 3 4 5

2

Clean fittings. Clean end of grease gun. Lubricate with proper amount and right type of lubricant. Lubricate within variance of frequency. Check runout on shafts and couplings. Check for soft foot. Align angular. Align horizontal. Align equipment specifications, not coupling specifications. Identify the proper tension and deflection for the belt. Set tension to specifications.

Consequences for not following best practices

Probability of future failures—number of self-induced failures vs. following best practices

Early bearing failure: reduced life by 20–80%.

100% 20 vs. 1

Premature coupling failure. Premature bearing and seal failure in motor and driven unit. Excessive energy loss.

100% 7 vs. 1

Premature belt failures through rapid belt wear or total belt failure. Premature bearing failure of driven and driver unit. Belt creeping or slipping causing speed variation without excessive noise. Motor shaft breakage.

100% 20 vs. 1

Continued

Table 1.1 continued

Maintenance task Hydraulic Components

Standard Hydraulic fluid must be conditioned to component specifications.

Required best practices 1

2

3

4

Hydraulic fluid must be input into the hydraulic reservoir utilizing a filter pumping system only. Filters must be rated to meet the needs of the component reliability and not equipment manufacturer’s specification. Filters must be changed on a timed basis based on filter condition. Oil samples must be taken on a set frequency, and all particles should be trended in order to understand the condition and wear of the hydraulic unit.

Consequences for not following best practices

Probability of future failures—number of self-induced failures vs. following best practices

Sticking hydraulic. Premature or 100% 30 vs. 1 unknown hydraulic pump life. Sustaining hydraulic competency by maintenance personnel. Length of equipment breakdown causes lost production.

Introduction: Why Use Best Maintenance Repair Practices? 5

MAINTENANCE The act of MAINTAINING.

MAINTAIN

KEEP in an existing state. PRESERVE from failure or decline. Keep Preserve Protect

Figure 1.1 Maintenance

In order to solve the problem of not following Best Maintenance Repair Practices, a sequential course of action should be taken: First, identify whether a problem exists (i.e., track repetitive equipment failures, review capacity losses in production and identify causes for these losses, and measure the financial losses due to repair issues). (See Figure 1.2, “The Maintenance Cost Iceberg.”) Second, identify the source of the problem (this could be combination of issues): ●

Maintenance skill level: Perform skills assessment (written and performance based) to evaluate whether skill levels are adequate to meet “Best Maintenance Repair Practices” for your specific maintenance organization.

●

Maintenance culture: Provide training to all maintenance and management relative to a change in maintenance strategy and how it will impact them individually (e.g., increase in profit for the plant, less overtime


Materials

Overtime

Contract Services Overhead & Benefits

Labor

Direct Maintenance Costs

Maintenance Budget Operating Budget

Indirect Maintenance Costs Set-Up

Ripple Effect On Production

Waste Disposal Start-Up Accidents Excess Inventory

Missed Schedules

Regulatory Fines

Environment Customer Complaints

Crisis Management Clean-Up

Lost Business

Product Liability

Emergency Purchases

Figure 1.2 The maintenance cost iceberg resulting from fewer equipment breakdowns, etc.). Track and measure the changes and display the results to everyone. ●

Maintenance strategy: Develop a plan to introduce a proactive maintenance model with “Preventive and Planned Maintenance” at the top of planned priorities. This will provide more time for performing maintenance utilizing the “Best Maintenance Repair Practices.”

Third, implement the changes needed to move toward following “Best Maintenance Repair Practices” and measure the financial gains. Everyone should be aware that financial rewards can be great but we must understand why they can also be hard to achieve. Several of the reasons why implementing a program of change, such as the one discussed, can be doomed to failure include: ●

Management not committed;

●

Lack of discipline and direction;


●

Lack of management commitment and accountability;

●

Momentum becomes slowed or changes direction;

●

Lack of an adequately skilled workforce;

●

No gap analysis or specific action plan to guide the effort to close the gaps;

●

Conflict between emergencies and performing maintenance following “Best Maintenance Repair Practices” (this does not mean all “emergent” repairs must be performed to “as built” specifications the first time, but it does mean that the repair, especially a temporary fix, will be corrected during the next outage of the equipment).

To conclude, as many as 90% of companies in the United States do not follow “Best Maintenance Repair Practices.” The percent that do follow these practices are realizing the rewards of a well run, capacity-driven organization that can successfully compete in today’s and tomorrow’s marketplace. Remember that use of the “Best Maintenance Repair Practices” might just become a mandatory requirement for the future success of an organization in today’s economy. Utilize this book as a resource to: 1

Write corrective maintenance procedures to attach to specific work order tasks in a computerized maintenance management system (CMMS).

2

Train personnel regarding new and existing maintenance repair procedures.

3

Be used as a tool by a current maintenance staff. Have all maintenance personnel use this book on the job site in order to follow “Best Maintenance Practices.” (This book can be set up in a storeroom in order to be replaced as the book is worn and damaged on the job site.)

Preventive and Predictive Maintenance (PPM) PPM is more than the regular cleaning, inspection, tightening, lubrication, and other actions intended to keep durable equipment in good operating condition and to avoid failures. It is an investment in the future—a future without major, disruptive breakdowns of critical equipment. It is,


at times, an investment without an immediate return on that investment. Here the philosophy must be, “Pay me now or pay me later,” because that is exactly what happens when it comes to PPM. If preventive maintenance is not accomplished in the proper way and in a timely manner, then the payme-later clause will occur at the most inopportune time. That is the premise that must be established and promoted by management. While the maintenance department has the day-to-day responsibility of plant PPM, the plant manager is ultimately responsible for setting the expectations concerning plant preventive maintenance. In general, the preventive and predictive maintenance effort is not focused. To spotlight some of the weaknesses, review the following points: 1

Many of the components that make up a good predictive maintenance program have not been developed;

2

Thermography;

3

Oil sampling for some gearboxes and hydraulic units.

Preventive Maintenance (PM) Some of the PM procedures have been developed; however, they lack details to make them efficient and safe, and to reinforce sound maintenance practices. 1

No mobile equipment has written PMs.

2

Logbooks are used, but it can’t be determined who was looking at and using the data.

3

All preventive maintenance regarding lubrication should be reviewed for detail and accuracy.

4

The forecasting and generation of PM tasks because of the state of the CMMS is not part of the normal maintenance routine.

Recommendations Review all lubrication-related task procedures for detail and content. ●

Provide training in effective plant lubrication procedures and techniques. The lubricator, maintenance personnel, and maintenance supervisors should attend.


●

The maintenance manager needs to conduct monthly spot checks on randomly selected PMs and repairs for quality assurance.

●

The present quantity and content of preventive maintenance tasks currently established are too generic in content. Task information should be detailed enough to help build consistency and training for maintenance— for example, “Line 1, Tire Stacker: On the full monthly PM, job #01 states: Lubricate, Check Photo Eyes, Clean/Sweep/Pickup, etc.” For the experienced maintenance person that has done the task many times, they most likely know what to look for. But the maintenance person with less experience does not. Again, provide enough information to build consistency and training.

●

Make the production associates aware of the fact that they are the eyes and ears of maintenance—the first line of defense—and that they are also an important part of the predictive maintenance process. They are the ones who will see or hear a problem first. Make sure to praise workers who contribute to the process.

●

Ensure all machines have operator checklists, that these checks are done properly, and that the results are turned in to the responsible supervisor.

●

Have the operators (in the future) perform routine clean-and-inspection tasks on equipment.

●

Encourage operator involvement concerning equipment PPM.

●

Train plant management and workforce in the importance of preventive and predictive maintenance.

Summary Preventive Maintenance is the most important routine function that maintenance personnel can accomplish. The reactive, breakdown maintenance mode will never be gotten away from if PMs are not performed consistently and properly on a regularly scheduled basis.

2


When most people think of preventive maintenance, they visualize scheduled, fixed interval maintenance that is done every day, every month, every quarter, every season, or at some other predetermined intervals. Timing may be based on days, or on intervals such as miles, gallons, activations, or hours of use. The use of performance intervals is itself a step toward basing preventive tasks on actual need, instead of just on a generality. The two main elements of fixed interval preventive maintenance are procedure and discipline. Procedure means that the correct tasks are done, the right lubricants applied, and consumables replaced at the best interval. Discipline requires that all the tasks are planned and controlled so that everything is done when it should be done. Both these areas deserve attention. The topic of procedures is covered in detail in following sections. Discipline is a major problem in many organizations. This is obvious when one considers the fact that many organizations do not have an established program. Further, organizations that do claim to have a program often fail to establish a good planning and control procedure to assure accomplishment. Elements of such a procedure include: 1

Listing of all equipment and the intervals at which it must receive PMs;

2

A master schedule for the year that breaks down tasks by month, week, and possibly even by the day;

3

Assignment of responsible persons to do the work;

4

Inspection by the responsible supervisor to make sure that quality work is done on time;

5

Updating of records to show when the work was done and when the next preventive task is due;

6

Follow-up as necessary to correct any discrepancies.

Fundamental Requirements of Effective Preventive/Predictive Maintenance 11

Fundamental Requirements of Effective Maintenance Effective maintenance is not magic, nor is it dependent on exotic technologies or expensive instruments or systems. Instead, it is dependent on doing simple, basic tasks that will result in reliable plant systems. These basics include:

Inspections Careful inspection, which can be done without “tearing down” the machine, saves both technician time and exposure of the equipment to possible damage. Rotating components find their own best relationship to surrounding components. For example, piston rings in an engine or compressor cylinder quickly wear to the cylinder wall configuration. If they are removed for inspection, chances are that they will not easily fit back into the same pattern. As a result, additional wear will occur, and the rings will have to be replaced much sooner than if they were left intact and performance-tested for pressure produced and metal particles in the lubricating oil.

Human Senses We humans have a great capability for sensing unusual sights, sounds, smells, tastes, vibrations, and touches. Every maintenance manager should make a concerted effort to increase the sensitivity of his own and that of his personnel’s human senses. Experience is generally the best teacher. Often, however, we experience things without knowing what we are experiencing. A few hours of training in what to look for could have high payoff. Human senses are able to detect large differences but are generally not sensitive to small changes. Time tends to have a dulling effect. Have you ever tried to determine if one color was the same as another without having a sample of each to compare side by side? If you have, you will understand the need for standards. A standard is any example that can be compared to the existing situation as a measurement. Quantitative specifications, photographs, recordings, and actual samples should be provided. The critical parameters should be clearly marked on them with displays as to what is good and what is bad. As the reliability-based preventive maintenance program develops, samples should be collected that will help to pinpoint with maximum accuracy

12 Fundamental Requirements of Effective Preventive/Predictive Maintenance

how much wear can take place before problems will occur. A display where craftsmen gather can be effective. A framed 4 × 4 pegboard works well since shafts, bearings, gears, and other components can be easily wired to it or hung on hooks for display. An effective, but little used, display area where notices can be posted is above the urinal or on the inside of the toilet stall door. Those are frequently viewed locations and allow people to make dual use of their time.

Sensors Since humans are not continually alert or sensitive to small changes and cannot get inside small spaces, especially when operating, it is necessary to use sensors that will measure conditions and transmit information to external indicators. Sensor technology is progressing rapidly; there have been considerable improvements in capability, accuracy, size, and cost. Pressure transducers, temperature thermocouples, electrical ammeters, revolution counters, and a liquid height level float are examples found in most automobiles. Accelerometers, eddy-current proximity sensors, and velocity seismic transducers are enabling the techniques of motion, position, and expansion analysis to be increasingly applied to large numbers of rotating equipment. Motors, turbines, compressors, jet engines, and generators can use vibration analysis. The normal pattern of operation, called its “signature,” is established by measuring the performance of equipment under known good conditions. Comparisons are made at routine intervals, such as every thirty days, to determine if any of the parameters are changing erratically, and further, what the effect of such changes may be. The spectrometric oil analysis process is useful for any mechanical moving device that uses oil for lubrication. It tests for the presence of metals, water, glycol, fuel dilution, viscosity, and solid particles. Automotive engines, compressors, and turbines all benefit from oil analysis. Most major oil companies will provide this service if you purchase lubricants from them. The major advantage of spectrometric oil analysis is early detection of component wear. Not only does it evaluate when oil is no longer lubricating properly and should be replaced, it also identifies and measures small quantities of metals that are wearing from the moving surfaces. The metallic elements found, and their quantity, can indicate what components are wearing and to what degree so that maintenance and overhaul can be carefully planned. For example, presence of chrome would indicate


cylinder-head wear, phosphor bronze would probably be from the main bearings, and stainless steel would point toward lifters. Experience with particular equipment naturally leads to improved diagnosis.

Thresholds Now that instrumentation is becoming available to measure equipment performance, it is still necessary to determine when that performance is “go” and when it is “no go.” A human must establish the threshold point, which can then be controlled by manual, semiautomatic, or automatic means. First, let’s decide how the threshold is set and then discuss how to control it. To set the threshold, one must gather information on what measurements can exist while equipment is running safely and what the measurements are just prior to or at the time of failure. Equipment manufacturers, and especially their experienced field representatives, will be a good starting source of information. Most manufacturers will run equipment until failure in their laboratories as part of their tests to evaluate quality, reliability, maintainability, and maintenance procedures. Such data are necessary to determine under actual operating conditions how much stress can be put on a device before it will break. Many devices, such as nuclear reactors and flying airplanes, should not be taken to the breaking point under operating conditions, but they can be made to fail under secure test conditions so that the knowledge can be used to keep them safe during actual use. Once the breaking point is determined, a margin of safety should be added to account for variations in individual components, environments, and operating conditions. Depending on the severity of failure, that safety margin could be anywhere from one to three standard deviations before the average failure point. One standard deviation on each side of the mean will include 68% of all variations, two standard deviations include 95%, and three standard deviations are 98.7%. Where our mission is to prevent failures, however, only the left half of the distribution is applicable. This singlesided distribution also shows that we are dealing with probabilities and risk. The earlier the threshold is set and effective preventive maintenance done, the greater is the assurance that it will be done prior to failure. If the mean time between failures (MTBF) is 9,000 miles with a standard deviation of 1,750 miles, then proper preventive maintenance at 5,500 miles


could eliminate almost 98% of the failures. Note the word “proper,” meaning that no new problems are injected. That also means, however, that costs will be higher than need be since components will be replaced before the end of their useful life, and more labor is required. Once the threshold set point has been determined, it should be monitored to detect when it is exceeded. The investment in monitoring depends on the period over which deterioration may occur, means of detection, and benefit value. If failure conditions build up quickly, a human may not easily detect the condition, and the relatively high cost of automatic instrumentation will be repaid.

Lubrication Friction of two materials moving relative to each other causes heat and wear. Friction-related problems cost industries over $1 billion per annum. Technology intended to improve wear resistance of metal, plastics, and other surfaces in motion has greatly improved over recent years, but planning, scheduling, and control of the lubricating program is often reminiscent of a plant handyman wandering around with his long-spouted oil can. Anything that is introduced onto or between moving surfaces in order to reduce friction is called a lubricant. Oils and greases are the most commonly used substances, although many other materials may be suitable. Other liquids and even gases are being used as lubricants. Air bearings, for example, are used in gyroscopes and other sensitive devices in which friction must be minimal. The functions of a lubricant are to: 1

Separate moving materials from each other in order to prevent wear, scoring, and seizure;

2

Reduce heat;

3

Keep out contaminants;

4

Protect against corrosion;

5

Wash away worn materials.

Good lubrication requires two conditions: sound technical design for lubrication and a management program to assure that every item of equipment is properly lubricated.


Lubrication Program Development Information for developing lubrication specifications can come from four main sources: 1

Equipment manufacturers;

2

Lubricant vendors;

3

Other equipment users;

4

Individuals’ own experience.

Like most other preventive maintenance elements, initial guidance on lubrication should come from manufacturers. They should have extensive experience with their own equipment both in their test laboratories and in customer locations. They should know what parts wear and are frequently replaced. Therein lies a caution: a manufacturer could, in fact, make short-term profits by selling large numbers of spare parts to replace worn ones. Over the long term, however, that strategy will backfire, and other vendors, whose equipment is less prone to wear and failure, will replace them. Lubricant suppliers can be a valuable source of information. Most major oil companies will invest considerable time and effort in evaluating their customers’ equipment to select the best lubricants and intervals for change. Naturally, these vendors hope that the consumer will purchase their lubricants, but the total result can be beneficial to everyone. Lubricant vendors perform a valuable service of communicating and applying knowledge gained from many users to their customers’ specific problems and opportunities. Experience gained under similar operating conditions by other users or in your own facilities can be one of the best teachers. Personnel, including operators and mechanics, have a major impact on lubrication programs. A major step in developing the lubrication program is to assign specific responsibility and authority for the lubrication program to a competent maintainability or maintenance engineer. The primary functions and steps involved in developing the program are to: 1

Identify every piece of equipment that requires lubrication;

2

Assure that all major equipment is uniquely identified, preferably with a prominently displayed number;


3

Assure that equipment records are complete for manufacturer and physical location;

4

Determine locations on each piece of equipment that needs to be lubricated;

5

Identify lubricant to be used;

6

Determine the best method of application;

7

Establish the frequency or interval of lubrication;

8

Determine if the equipment can be safely lubricated while operating, or if it must be shut down;

9

Decide who should be responsible for any human involvement;

10 Standardize lubrication methods; 11 Package the above elements into a lubrication program; 12 Establish storage and handling procedures; 13 Evaluate new lubricants to take advantage of state of the art; 14 Analyze any failures involving lubrication and initiate necessary corrective actions. An individual supervisor in the maintenance department should be assigned the responsibility for implementation and continued operation of the lubrication program. This person’s primary functions are to: 1

Establish lubrication service actions and schedules;

2

Define the lubrication routes by building, area, and organization;

3

Assign responsibilities to specific persons;

4

Train lubricators;

5

Assure supplies of proper lubricants through the storeroom;

6

Establish feedback that assures completion of assigned lubrication and follows up on any discrepancies;

7

Develop a manual or computerized lubrication scheduling and control system as part of the larger maintenance management program;


8

Motivate lubrication personnel to check equipment for other problems and to create work requests where feasible;

9

Assure continued operation of the lubrication system.

It is important that a responsible person who recognizes the value of thorough lubrication be placed in charge. As with any activity, interest diminishes over time, equipment is modified without corresponding changes to the lubrication procedures, and state-of-the-art advances in lubricating technology may not be undertaken. A factory may have thousands of lubricating points that require attention. Lubrication is no less important to computer systems, even though they are often perceived as electronic. The computer field engineer must provide proper lubrication to printers, tape drives, and disks that spin at 3,600 rpm. A lot of maintenance time is invested in lubrication. The effect on production uptime can be measured nationally in billions of dollars.

Calibration Calibration is a special form of preventive maintenance whose objective is to keep measurement and control instruments within specified limits. A “standard” must be used to calibrate the equipment. Standards are derived from parameters established by the National Bureau of Standards (NBS). Secondary standards that have been manufactured to close tolerances and set against the primary standard are available through many test and calibration laboratories and often in industrial and university tool rooms and research labs. Ohmmeters are examples of equipment that should be calibrated at least once a year and before further use if subjected to sudden shock or stress. The government sets forth calibration system requirements in MIL-C-45662 and provides a good outline in the military standardization handbook MIL-HDBK-52, Evaluation of Contractor’s Calibration System. The principles are equally applicable to any industrial or commercial situation. The purpose of a calibration system is to provide for the prevention of tool inaccuracy through prompt detection of deficiencies and timely application of corrective action. Every organization should prepare a written description of its calibration system. This description should cover the measuring of test equipment and standards, and should: 1

Establish realistic calibration intervals;

2

List all measurement standards;


3

List all environmental conditions for calibration;

4

Ensure the use of calibration procedures for all equipment and standards;

5

Coordinate the calibration system with all users;

6

Assure that equipment is frequently checked by periodic system or cross-checks in order to detect damage, inoperative instruments, erratic readings, and other performance degrading factors that cannot be anticipated or provided for by calibration intervals;

7

Provide for timely and positive correction action;

8

Establish decals, reject tags, and records for calibration labeling;

9

Maintain formal records to assure proper controls.

The checking interval may be in terms of time—hourly, weekly, monthly— or based on amount of use—every 5,000 parts, or every lot. For electrical test equipment, the power-on time may be the critical factor and can be measured through an electrical elapsed-time indicator. Adherence to the checking schedule makes or breaks the system. The interval should be based on stability, purpose, and degree of usage. If initial records indicate that the equipment remains within the required accuracy for successive calibrations, then the intervals may be lengthened. On the other hand, if equipment requires frequent adjustment or repair, the intervals should be shortened. Any equipment that does not have specific calibration intervals should be (1) examined at least every six months, and (2) calibrated at intervals of no longer than one year. Adjustments or assignment of calibration intervals should be done in such a way that a minimum of 95% of equipment, or standards of the same type, is within tolerance when submitted for regularly scheduled recalibration. In other words, if more than 5% of a particular type of equipment is out of tolerance at the end of its interval, then the interval should be reduced until less than 5% is defective when checked. A record system should be kept on every instrument, including: 1

History of use;

2

Accuracy;

3

Present location;


4

Calibration interval and when due;

5

Calibration procedures and necessary controls;

6

Actual values of latest calibration;

7

History of maintenance and repairs.

Test equipment and measurement standards should be labeled to indicate the date of last calibration, by whom it was calibrated, and when the next calibration is due. When the size of the equipment limits the application of labels, an identifying code should be applied to reflect the serviceability and due date for the next calibration. This provides a visual indication of the calibration serviceability status. Both the headquarters calibration organization and the instrument user should maintain a two-way check on calibration. A simple means of doing this is to have a small form for each instrument with a calendar of weeks or months (depending on the interval required) across the top, which can be punched and noticed to indicate the calibration due date.

Planning and Estimating Planning is the heart of good inspection and preventive maintenance. As described earlier, the first thing to establish is what items must be maintained and what the best procedure is for performing that task. Establishing good procedures requires a good deal of time and talent. This can be a good activity for a new graduate engineer, perhaps as part of a training process that rotates him or her through various disciplines in a plant or field organization. This experience can be excellent training for a future design engineer. Writing ability is an important qualification, along with pragmatic experience in maintenance practices. The language used should be clear and concise, using short sentences. Who, what, when, where, why, and how should be clearly described. The following points should be noted from this typical procedure: 1

Every procedure has an identifying number and title;

2

The purpose is outlined;

3

Tools, reference documents, and any parts are listed;

4

Safety and operating cautions are prominently displayed;

5

A location is clearly provided for the maintenance mechanic to indicate performance as either satisfactory or deficient. If deficient,


details are written in the space provided at the bottom for planning further work. The procedure may be printed on a reusable, plastic-covered card that can be pulled from the file, marked, and returned when the work order is complete; on a standard preprinted form; or on a form that is uniquely printed by computer each time a related work order is prepared. Whatever the medium of the form, it should be given to the preventive maintenance craftsperson together with the work order so that he has all the necessary information at his fingertips. The computer version has the advantage of single-point control that may be uniformly distributed to many locations. This makes it easy for an engineer at headquarters to prepare a new procedure or to make any changes directly on the computer and have them instantly available to any user in the latest version. Two slightly different philosophies exist for accomplishing the unscheduled actions that are necessary to repair defects found during inspection and preventive maintenance. One is to fix them on the spot. The other is to identify them clearly for later corrective action. If a “priority one” defect that could hurt a person or cause severe damage is observed, the equipment should be immediately stopped and “46 red tagged” so that it will not be used until repairs are made. Maintenance management should establish a guideline such as, “Fix anything that can be corrected within ten minutes, but if it will take longer, write a separate work request.” The policy time limit should be set based on: 1

Travel time to that work location;

2

Effect on production;

3

Need to keep the craftsperson on a precise time schedule.

The inspector who finds them can effect many small repairs the most quickly. This avoids the need for someone else to travel to that location, identify the problem, and correct it. And it provides immediate customer satisfaction. More time-consuming repairs would disrupt the inspector’s plans, which could cause other, even more serious problems to go undetected. The inspector is like a general practitioner, who performs a physical exam and may give advice on proper diet and exercise but who refers any problems he may find to a specialist. The inspection or preventive maintenance procedure form should have space where any additional action required can be indicated. When the


procedure is completed and turned into maintenance control, the planner or scheduler should note any additional work required and see that it gets done according to priority.

Estimating Time Since inspection or preventive maintenance is a standardized procedure with little variation, the tasks and time required can be accurately estimated. Methods of developing time estimates include consideration of such resources as: 1

Equipment manufacturers’ recommendations;

2

National standards such as Chilton’s on automotive or Means’ for facilities;

3

Industrial engineering time-and-motion studies;

4

Historical experience.

Experience is the best teacher, but it must be carefully critiqued to make sure that the “one best way” is being used and that the pace of work is reasonable. The challenge in estimating is to plan a large percentage of the work (preferably at least 90%) so that the time constraints are challenging but achievable without a compromise in high quality. The trade-off between reasonable time and quality requires continuous surveillance by experienced supervisors. Naturally, if a maintenance mechanic knows that his work is being time studied, he will follow every procedure specifically and will methodically check off each step of the procedure. When the industrial engineer goes away, the mechanic will do what he feels are necessary items, in an order that may or may not be satisfactory. As discussed earlier in regard to motivation, an experienced preventive maintenance inspector mechanic can vary performance as much as 50% either way from the standard without most maintenance supervisors recognizing a problem or opportunity for improvement. Periodic checking against national or time-and-motion standards, as well as trend analysis of repetitive tasks, will help keep preventive task times at a high level of effectiveness.

Estimating Labor Cost Cost estimates follow from time estimates simply by multiplying the hours required by the required labor rates. Beware of coordination problems where multiple crafts are involved. For example, one Fortune 100 company


has trade jurisdictions that require the following personnel in order to remove an electric motor: a tinsmith to remove the cover, an electrician to disconnect the electrical supply, a millwright to unbolt the mounts, and one or more laborers to remove the motor from its mound. That situation is fraught with inefficiency and high labor costs, since all four trades must be scheduled together, with at least three people watching while the fourth is at work. The cost will be at least four times what it could be, and is often greater if one of the trades does not show up on time. The best a scheduler can hope for is if he has the latitude to schedule the cover removal at, say, 8:00 A.M., and the other functions at reasonable time intervals thereafter: electrician at 9:00, millwright at 10:00, and laborers at 11:00. It is recommended that estimates be prepared on “pure” time. In other words, the exact hours and minutes that would be required under perfect scheduling conditions should be used. Likewise, it should be assumed that equipment would be immediately available from production. Delay time should be reported, and scheduling problems should be identified so that they can be addressed separately from the hands-on procedure times. Note that people think in hours and minutes, so one hour and ten minutes is easier to understand than 1.17 hours.

Estimating Materials Most parts and materials that are used for preventive maintenance are well known and can be identified in advance. The quantity of each item planned should be multiplied by the cost of the item in inventory. The sum of those extended costs will be the material cost estimate. Consumables such as transmission oil should be enumerated as direct costs, but grease and other supplies used from bulk should be included in overhead costs.

Scheduling Scheduling is, of course, one of the advantages to doing preventive maintenance over waiting until equipment fails and then doing emergency repairs. Like many other activities, the watchword should be “PADA,” which stands for “Plan a Day Ahead.” In fact, the planning for inspections and preventive activities can be done days, weeks, and even months in advance in order to assure that the most convenient time for production is chosen, that maintenance parts and materials are available, and that the maintenance workload is relatively uniform.


Scheduling is primarily concerned with balancing demand and supply. Demand comes from the equipment’s need for preventive maintenance. Supply is the availability of the equipment, craftspeople, and materials that are necessary to do the work. Establishing the demand is partially covered in the chapters on on-condition, condition monitoring, and fixed interval preventive maintenance tasks. Those techniques identify individual equipment as candidates for PM.

Coordination with Production Equipment is not always available for preventive maintenance just when the maintenance schedulers would like it to be. An overriding influence on coordination should be a cooperative attitude between production and maintenance. This is best achieved by a meeting between the maintenance manager and production management, including the foreman level, so that what will be done to prevent failures, how this will be accomplished, and what production should expect to gain in uptime may all be explained. The cooperation of the individual machine operators is of prime importance. They are on the spot and most able to detect unusual events that may indicate equipment malfunctions. Once an attitude of general cooperation is established, coordination should be refined to monthly, weekly, daily, and possibly even hourly schedules. Major shutdowns and holidays should be carefully planned so any work that requires “cold” shutdown can be done during those periods. Maintenance will often find that they must do this kind of work on weekends and holidays, when other persons are on vacation. Normal maintenance should be coordinated according to the following considerations: 1

Maintenance should publish a list of all equipment that is needed for inspections, preventive maintenance, and modifications, and the amount of cycle time that such equipment will be required from production.

2

A maintenance planner should negotiate the schedule with production planning so that a balanced workload is available each week.

3

By Wednesday of each week, the schedule for the following week should be negotiated and posted where it is available to all concerned; it should be broken down by days.

4

By the end of the day before the preventive activity is scheduled, the maintenance person who will do the PM should have seen the first-line


production supervisor in charge of the equipment to establish a specific time for the preventive task. 5

The craftsperson should make every effort to do the job according to schedule.

6

As soon as the work is complete, the maintenance person should notify the production supervisor so that the equipment may be put back into use.

Overdue work should be tracked and brought up-to-date. Preventive maintenance scheduling should make sure that the interval is maintained between preventive actions. For example, if a preventive task for May is done on the thirtieth of the month, the next monthly task should be done during the last week of June. It is foolish to do a preventive maintenance task on May 30th and another June 1st, just to be able to say one was done each month. In the case of preventive maintenance, the important thing is not the score but how the game was played.

Assuring Completion A formal record is desirable for every inspection and preventive maintenance job. If the work is at all detailed, a checklist should be used. The completed checklist should be returned to the maintenance office on completion of the work. Any open preventive maintenance work orders should be kept on report until the supervisor has checked the results for quality assurance and signed off approval. Modern computer technology with handheld computers and pen-based electronic assistants permits paperless checklists and verification. In many situations, a paper work order form is still the most practical media for the field technician. The collected data should then be entered into a computer system for tracking.

Record Keeping The foundation records for preventive maintenance are the equipment files. In a small operation with less than 200 pieces of complex equipment, the records can easily be maintained on paper. The equipment records provide information for purposes other than preventive maintenance. The essential items include: ●

Equipment identification number;

●

Equipment name;


●

Equipment product/group/class;

●

Location;

●

Use meter reading;

●

PM interval(s)

●

Use per day;

●

Last PM due;

●

Next PM due;

●

Cycle time for PM;

●

Crafts required, number of persons, and time for each;

●

Parts required.

Back to Basics Obviously, effective maintenance management requires much more than these fundamental tasks. However, these basic tasks must be the foundation of every successful maintenance program. The addition of other tools, such as CMMS, predictive maintenance, etc., cannot replace them.

3


Introduction A maintenance skills assessment is a valuable tool in determining the strengths and weaknesses of an individual or a given group of employees in order to design a high-impact training program that targets those documented needs. Maintenance personnel have often found it difficult to upgrade their technical skills because much that is available is redundant or does not take their current skill level into consideration. An assessment is designed to eliminate those problems by facilitating the construction of customized training paths for either individuals or the group based upon demonstrated existing knowledge and skills. When used in conjunction with a job task analysis, a gap analysis can be performed to determine both what skills are needed in order to perform the job effectively and what skills those in the workforce presently have.

Definition of a Skills Assessment A maintenance skills assessment consists of a series of written tests, performance exercises, and identification activities selected from a listing of mechanical basic skill areas. In this chapter maintenance mechanics will be able to assess their maintenance knowledge only because skills can only be assessed through a hands-on assessment. However, the knowledge assessment is the prerequisite for all skills. The written assessment in this chapter is written at an eighth-grade level (maintenance mechanics, in most industries, must be able to read proficiently at least at the 12-year twelfth-grade level). A maintenance person without the knowledge required for a specific skill can be assured mistakes will be made in mechanical judgment and ability and thus will cause equipment failures. This knowledge assessment will not cover all skill areas covered in this book but will cover chapters that are considered the mechanical basics.

Maintenance Skills Assessment 27

Knowledge Assessment This knowledge assessment is directed to the following skills. The answers will be provided in Appendix A at the end of the book. A minimum score of 90% in each skill area should be considered acceptable in most cases. However, some skill areas could require a higher score if the risk of failure due to a knowledge gap is high. In all areas of maintenance, a maintenance person must know the risk. The knowledge assessment will be performed in the following skills areas: ●

Safety

●

Lubrication

●

Bearings

●

Chain Drives

●

Belt Drives

●

Hydraulics

●

Couplings

Knowledge Assessment The assessment is multiple choice. Select the best answer. Do not look at the answers until you have answered all the questions.

Knowledge Area: Safety 1

What term is used to describe places where moving parts meet or come near each other? A. Guard B. Closer

28 Maintenance Skills Assessment

C. Pinch points D. Assembly point 2

What is another name for back-and-forth motion? A. B. C. D.

3

Reciprocating motion Away Advancing lateral None of the above

What is the term “point of operation?” A. The section of the process where the machine centers meet B. The main focus of process C. The place where the raw material or work-piece is processed by a machine D. A point where operators cannot see

4

If a bench grinder is equipped with safety guards, is it necessary for you to wear personal protective equipment? A. Yes B. No

5

What type of machine guard limits the operator’s access to the danger zone? A. B. C. D.

6

What type of machine guard prevents access to the danger zone altogether? A. B. C. D.

7

Safety chain E-stop A barrier guard None of the above

An enclosure guard Safety chain Barrier guard None of the above

What kind of guards cannot be moved when a machine is in operation? A. B. C. D.

Barrier guards Fixed guards E-stop guards None of the above


8

What type of guard prevents a machine from operating when the guard is opened or removed? A. B. C. D.

9

Barrier guard Tapeless guard An interlocking guard None of the above

What type machine guard is capable of physically pulling an operator’s hands out of the danger zone? A. B. C. D.

An automatic guard Barrier guard Restrictive guard None of the above

10 When will a presence-sensing guard stop a machine? A. B. C. D.

When a person is located outside of the danger zone When a light signals a safety alert There is no such item When a person or object enters the danger zone

11 What kind of controls does a machine have if the operator must remove both hands from the danger zone in order to start the machine? A. B. C. D.

Hand-free controls Two-hand trip controls Standard controls Automatic controls

12 What kinds of tools make it unnecessary for an operator to reach into the danger zone? A. B. C. D.

Hand-extraction tools Local guarding tools Feeding and extracting tools None of the above

13 Under what conditions would you remove someone else’s lock from a lockout device? A. When plant manager or maintenance manager approves B. When the person that installed the lockout cannot be found after 30 minutes C. When you think it is OK to do so safely D. According to your plant’s lockout procedure


14 What must your employer provide in addition to the appropriate PPE? A. B. C. D.

Training in its use Safety bulletins Training material and trainers None of the above

15 What is your responsibility before using PPE? A. B. C. D.

None You must inspect it Locate all documents controlling it You must report its condition to your supervisor

16 Why should you avoid loose-fitting clothing in the plant? A. B. C. D.

It can create a barrier from sharp parts It can get caught in moving parts It is unprofessional None of the above

17 What should you do if you accidentally come in contact with a dangerous chemical? A. B. C. D.

It depends on the chemical Report to the safety director Shower for at least 15 minutes to rinse thoroughly None of the above

18 How much clearance should hard-hat webbing provide between your head and the top of the shell? A. B. C. D.

Close as possible 1" 2 Does not matter 1"

19 How can you keep dust and liquids from reaching your eyes from any direction? A. B. C. D.

By wearing safety goggles By wearing safety glasses By wearing a face shield Any of the above


20 What units are used to measure noise? A. B. C. D.

Trebels Decibels Milibars None of the above

Knowledge Area: Lubrication 1

A lubricant’s viscosity is rated by what type of unit? A. B. C. D.

2

A lubricant with high viscosity has a: A. B. C. D.

3

Provides good cushioning for machine shock loads. Can flow into tight spaces for better lubrication. Does not carry heat away as well as a high-viscosity lubricant. Costs less than a high-viscosity lubricant.

What are two disadvantages of high-viscosity lubricants? A. B. C. D.

5

High speed. High temperature. High resistance to flow. High resistance to breakdown.

A low-viscosity lubricant: A. B. C. D.

4

SSU SAE ISA LVU

They are expensive and cannot be used on high-speed motors. They break down quickly and are difficult to apply. They do not flow well and do not carry heat away well. They do not protect against abrasive action of dirt, and they break down quickly.

Multiple-viscosity lubricants differ from single-viscosity lubricants because: A. They have special additives that extend their effective temperature range. B. They are best within a very narrow temperature range.


C. They will never degrade under high temperatures. D. They last longer. 6

One advantage of multiple-viscosity lubricants is that: A. They flow better at medium range temperatures than at either extreme. B. They have a high bearing capacity. C. They have a broad working temperature range. D. They do not break down in the presence of water.

7

Which of the following is NOT a factor affecting the selection of a lubricant? A. B. C. D.

8

When choosing a lubricant, you want: A. B. C. D.

9

Machine speed Environmental humidity Operating temperature Environmental temperatures

The lubricant to stay thin at high temperatures. The lubricant to thicken at low temperatures. The lubricant to thin at low temperatures. The lubricant to maintain effective viscosity at its highest rated temperatures.

An oil cooler is used to: A. Add heat to the oil to enable it to flow better at low temperatures. B. Add heat to the oil to keep it from thinning at high temperatures. C. Remove heat from the oil to prevent it from thinning at high temperatures. D. Remove heat from the oil to prevent it from thickening at low temperatures.

10 What function do detergent additives in lubricants perform? A. B. C. D.

Keep metal surfaces clean Keep the lubricant clean Minimize the amount of foaming All of the above

11 An anti-oxidation additive in a lubricant: A. Controls the level of dirt. B. Controls the amount of mixing with air.


C. Controls the level of foaming. D. Prevents the lubricant from mixing with metal particles. 12 As a mechanic, you observe that a machine bearing is extremely hot and becoming discolored as it operates. Your conclusion is that the: A. B. C. D.

Lubricant is contaminated by water. Bearing is about to seize. Lubricant is causing acid corrosion on the bearing. Bearing is not compatible with the lubricant.

13 When cooling an overheated bearing, what should you do first? A. B. C. D.

Wrap the bearing housing in hot, wet rags. Spray cool water on the bearing. Inject cool oil in the bearing. Wrap the bearing housing in cool, wet rags.

14 Oil returning to the sump is visually cloudy and foaming. You conclude that the oil is: A. B. C. D.

Contaminated with soot. Contaminated with water. Contaminated with metal particles. In need of detergent additives.

15 Undesired oil misting can be reduced by: A. B. C. D.

Increasing the temperature of the oil. Increasing the speed of the machine. Increasing the viscosity of the oil. Reducing the viscosity of the oil.

16 A grease cup is defined as a: A. Cup filled with grease that screws onto a fitting. B. Timed lubrication system controlled by a rotating cam. C. Gravity system that forces lubricant onto or into the area needing lubrication. D. Fitting that applies oil in droplet form. 17 A lubricating system used in low-speed applications in which a needle valve meters a steady rate of lubricant to a machine without recycling the lubricant is a(n): A. Dip lubricator. B. Shot lubricator.


C. Drip lubricator. D. Oil sump. 18 A lubrication system in which the component needing lubrication rotates through an enclosed housing containing oil and carries the oil to other components is called a(n): A. B. C. D.

Dip lubricator. Shot lubricator. Drip lubricator. Oil sump.

19 In a force-feed lubrication system, lubricant is moved to the component needing lubrication by a: A. B. C. D.

Cooler. Pump. Filter. Bearing.

20 What is the most undesirable by-product of oil misting? A. B. C. D.

Bearing failure Shaft damage Explosion potential Oil breakdown

Knowledge Area: Bearings 1

The two basic categories of bearings are: A. B. C. D.

2

Plain and antifriction. Ball and roller. Journal and ball. Pillow-block and roller.

Bearings: A. B. C. D.

Are found in machines with moving parts. Function as guides. Help reduce the friction between moving parts. All of the above.


3

Thrust bearings: A. B. C. D.

4

Antifriction bearings: A. B. C. D.

5

Separating the races with balls or rollers. Using an air gap. Hydraulics. Running on a film of lubricant.

Roller bearings are used over ball bearings for which of the following situations? A. B. C. D.

9

Separating the races with balls or rollers. Using an air gap. Hydraulics. Running on a film of lubricant.

Antifriction bearings operate by: A. B. C. D.

8

Lubrication by hand. Central grease systems. Pressure-feed oil systems. All of the above.

Plain bearings operate by: A. B. C. D.

7

Contain balls. Contain rollers. Will run hot if they are overlubricated. All of the above.

Bearing lubrication systems include: A. B. C. D.

6

Support axial loads on rotating members. Support radial loads on rotating members. Both A and B. None of the above.

High-speed applications High-load applications Wet environments Mobile equipment engines

Bearing clearance can be described as: A. The space between the rolling elements and the races. B. The allowed difference between the shaft size and the bearing inner race. C. The allowed differences between bearing inner and outer race. D. None of the above.


10 Shaft tolerance can be defined as: A. The allowed difference between the shaft size and the bearing inner race. B. The force applied during installation. C. The space between the rolling elements and the races. D. None of the above. 11 The preferred method for installing an antifriction bearing is: A. B. C. D.

With a sledge hammer. To sand down the shaft until the bearing slides on. With a bearing heater. Both B and C.

12 When tightening the lock-nut on a spherical roller bearing, the preferred tool is: A. B. C. D.

A spanner wrench. A bearing heater. A hammer and punch. None of the above.

13 The bearing best suited for both radial and thrust loads is a bearing. A. B. C. D.

tapered-sleeve linear-motion needle tapered-roller

14 A bearing lubricated with oil is capable of bearing lubricated with grease. A. B. C. D.

speeds than the same

lower higher the same different

15 As you tighten the nut on a spherical-roller bearing, the space between the race and the rolling element: A. B. C. D.

Increases. Decreases. Remains the same. Develops cracks.


16 On a metric bearing with the number 7307, the ID of the bearing is: A. B. C. D.

35 mm. 7 mm. .035". .007".

17 To convert the metric shaft size of a bearing to inches, you multiply the millimeters by: A. B. C. D.

5. 39. .03937. .05.

18 A failed bearing that has a cracked inner race probably failed because: A. B. C. D.

the shaft was too large. of a lack of lubricant. the operator failed to do the proper inspection. of overlubrication.

19 An antifriction bearing can run hot because: A. B. C. D.

of overlubrication. it is about to fail. of excessive load. all of the above.

20 A 20% increase in bearing load, can result in a bearing life. A. B. C. D.

% decrease in

20 100 50 10

Knowledge Area: Chain Drives 1

What is the maximum elongation that a roller chain can tolerate before it should be replaced? A. 10% B. 3 inches


C. 3% 3 inch per ft. of chain D. 16 2

What causes roller fatigue? A. B. C. D.

3

What problems will chain misalignment cause? A. B. C. D.

4

Safety chain E-stop A barrier guard None of the above

What type of machine guard prevents access to the danger zone altogether? A. B. C. D.

7

Lubrication and alignment Lubrication and tension Tension and alignment All of the above

What type of machine guard limits the operator’s access to the danger zone? A. B. C. D.

6

Side plate wear Sprocket wear Pin wear Chain break

What are the two most important factors in the life of a silent chain drive? A. B. C. D.

5

Lack of lubrication Contamination Improper chain tension All of the above


Name the two main groups of mechanical couplings: A. B. C. D.

Rigid and flexible. Chain and gear. Grid and Lovejoy. Spring and centrifugal.


8

Large fans are typically powered in what way? A. B. C. D.

9

Motor and belts Motor and chain drive Motor direct drive Motor, gearbox, and belts

Which type of conductivity cells are usually equipped with a safety chain to prevent the cell from blowing out? A. B. C. D.

Immersions Insertion Flow Screw-in

10 Chain-linked fences should be topped with: A. Three strands of electrically energized wire. B. Barbed wire. C. A skirt of the same material that kicks back 45 degrees and extends 18 inches. D. None of the above. 11 What are the two links that make up a standard roller chain? A. B. C. D.

Roller, rigid Roller, pin Master, half Offset, half

12 On a roller chain with a designation of 35, what do the 3 and the 5 designate? A. B. C. D.

3 is the size in inches; 5 is the duty 3 is the duty; 5 is the size in thousandths of an inch 3 is the size in eighths of an inch; 5 means a rollerless chain 3 is the size; 5 means heavy duty

13 What is the maximum elongation that a roller chain can tolerate before it should be replaced? A. 10% B. 3 inches C. 3% 3 inch per ft. of chain D. 16


14 What causes roller fatigue? A. Lack of lubrication B. Contamination C. Improper chain tension D. All of the above 15 What problems will chain misalignment cause? A. Side plate wear B. Pin wear C. Chain break D. All of the above 16 What are the two most important factors in the life of a silent chain drive? A. Lubrication and alignment B. Lubrication and tension C. Tension and alignment D. Tension and speed 17 Name the two main groups of mechanical couplings. A. Rigid and flexible B. Chain and gear C. Grid and Lovejoy D. Spring and centrifugal 18 What type of machine guard limits the operator’s access to the danger zone? A. Safety chain B. E-stop C. A barrier guard D. None of the above 19 What type of machine guard prevents access to the danger zone altogether? A. B. C. D.


20 What is the tension or slack a roller chain is recommended to have? A. B. C. D.

10% slack between centers 1 inch slack 2 5% slack between centers 2% slack between centers


Knowledge Area: Belt Drives 1

Of the following, what are the four main types of belt drives? A. B. C. D.

2

Identify routine maintenance performed on flat belts. A. B. C. D.

3

They do not require tensioning. The material they are made of reduces slippage. They provide slip-proof engagement. There is not load fluctuation.

Different speeds are required for various materials when cutting with a band saw. How is saw speed adjusted? A. B. C. D.

7

Oil, dirt, heat, and alignment Water, dirt, sunlight, and tension Guarding, shielding, reflection, and direction Sunlight, dry-rot, mold, and water

Why do timing or gear belts not require high tension? A. B. C. D.

6

Reduce center-to-center distance of the pulleys. Use a pry bar to slide the belt over the pulleys. Tension the belt until your hand bounces off the belt. All of the above.

Which of the following contribute to rapid belt wear? A. B. C. D.

5

Tensioning Cleaning Dressing Both A and B

Which of the following is an acceptable V-belt installation process? A. B. C. D.

4

V-belt; flat belt; timing belt; ribbed belt Straight belt; V-belt; gear belt; ribbed belt Rubber; vitron; buna-n; teflon Round belt; dual face belt; V-belt; multibelt

Set the speed indicator to the desired rpm and start the machine. Adjust the belt tensioner and start the machine. Start the machine and adjust the speed indicator. Dial in the correct job selector and start the machine.

Belt drives are designed to transmit power between a drive unit and a: A. Conveyor belt. B. Motor.


C. Driven unit. D. Roller. 8

A drive belt with the designation “XL” would indicate which of the following types? A. B. C. D.

9

Cogged, fractional horsepower Conventional V-belt Narrow V-belt Positive drive belt

A drive belt with the designation “C” would indicate which of the following types? A. B. C. D.

Power band belt Conventional V-belt Narrow V-belt Fractional horsepower

10 A drive belt with the designation “V” would indicate which of the following types? A. B. C. D.

Positive drive belt Conventional V-belt Narrow V-belt Fractional horsepower

4 " long. What 11 A “C” belt with a nominal length of 60 inches is 10 identification should be marked on the belt?

A. B. C. D.

C410 C60/4 C64 C56

12 When replacing one of a set of belts, the mechanic should: A. B. C. D.

Replace the worn belt only. Replace the worn belt and the belt closest to the drive motor. Replace all of the belts. Replace the worn belt and the belt farthest from the drive motor.

13 What criteria should be used to determine if a belt needs to be replaced? 1 ". A. The belt protrudes above the top of the sheave no more than 16 B. The belt is flush with the top of the sheave groove. 1 " into the groove sheave. C. The belt is recessed more than 16

D. The belt is recessed more than 18 " into the groove sheave.


14 The length of belt is determined by its: A. Pitch line. B. V-line. C. Standard line. D. None of the above. 15 When “timing” is a critical consideration with a belt, which of the following is typically used? A. Joined belt B. V-belt C. Cogged belt D. Positive drive belt 16 Regarding sheave grove angle, which of the following statements is correct? A. Smaller diameter sheaves have less groove angle than larger diameter grooves. B. Larger diameter sheaves have less groove angle than smaller diameter sheaves. C. Groove angles are not affected by sheave diameter. D. Can be compensated for by adjusting belt tension. 17 Which of the following conditions can cause excessive belt wear? A. B. C. D.

Dirty operating conditions Improper sheave alignment Improper tensioning All of the above

Knowledge Area: Hydraulics 1

What function does a reservoir serve in a hydraulic system? A. B. C. D.

2

Stores surplus of oil Cools the oil Cleans the oil A, B, and C

Why is there a baffle in a hydraulic reservoir? A. B. C. D.

To prevent baffling To assist in cleaning and cooling of the oil To keep the oil level the same To prevent overpressurization of the reservoir


3

What should the fluid pass through before it reaches the pump? A. B. C. D.

4

Prepump precipitator Metal diffuser Heat exchanger Strainer

What are the three basic types of positive displacement pumps? A. Centrifugal, gear, piston B. Air pump, piston, centrifugal C. Piston, gear, vane

5

What function does a relief valve provide in a hydraulic system? A. B. C. D.

6

What function does an actuator serve in a hydraulic system? A. B. C. D.

7

Open, close, tandem, return Bidirectional, unidirectional, terminal, return Tandem, float, closed, open The center position is the same in all valves

What function do flow control valves serve in the hydraulic system? A. B. C. D.

9

To hold pressure in a system To assist the pump in flow To allow work to be performed A and B

What are the four basic types of center conditions (envelopes) in the directional control valves of a hydraulic system? A. B. C. D.

8

Protects the system from overpressurization Dumps oil back to the tank Cools the oil A and B

Control pressure to a point in the system. Control the speed of an actuator. Reduce the speed of a pump. Eliminate return in one direction.

What are the two basic types of flow control arrangements in a hydraulic system? A. Meter in, meter out B. Flow control, flow decontrol


C. Meter in, direct flow D. Control, directional control 10 What are the two basic types of actuators in a hydraulic system? A. B. C. D.

Bladder, piston Cylinder, motor Double cylinder, single cylinder Pump, actuator

11 Which of the following are means by which a directional control valve may be actuated? A. B. C. D.

Manual Solenoid operated Pilot pressure All of the above

12 The variable volume pump used most often in high-pressure systems is what type of pump? A. B. C. D.

Vane Piston Gear Centrifugal

13 Basic hydraulics theory tells us that pressure: A. On a confined liquid is transmitted undiminished in every direction. B. Acts with equal force on equal areas. C. Acts at right angles only on the container side walls. D. Acts with equal force on reduced areas. 14 Hydraulics is the science of transmitting force or motion through a medium of a: A. Confined liquid. B. Nonconfined liquid. C. A and B. D. None of the above. 15 Component parts connected to each other by fluid lines in the hydraulic system form a: A. B. C. D.

Closed center system. Bypass. Circuit. Motor.


16 A needle valve in a hydraulic line provides a: A. B. C. D.

Flow control orifice. Pressure control orifice. Relief valve. Master valve.

Knowledge Area: Couplings 1

A major advantage of flexible couplings is that they: A. B. C. D.

2

Name the two main groups of mechanical couplings: A. B. C. D.

3

Slider Steelflex Torus rubber Sure-Flex rubber

Measurements taken at the top and bottom of a coupling are taken to adjust: A. B. C. D.

5

Rigid and flexible. Chain and gear. Grid and Lovejoy. Spring and centrifugal.

Which of the following couplings requires lubrication? A. B. C. D.

4

Compensate for thermal expansion. Provide greater torsional transfer than rigid couplings. Require only minimal motor alignment. Operate at higher rpm than rigid couplings.

Vertical alignment. Horizontal alignment. Angular alignment. Parallel alignment.

When performing coupling alignment, the first activity that must be conducted is to: A. B. C. D.

Check runout. Verify angular misalignment. Verify parallel misalignment. Check coupling spacing.


6

Which of the following measuring tools is most useful for the straightedge method of coupling alignment? A. B. C. D.

7

Cooling tower fan shafts are supported by: A. B. C. D.

8

Hubs. Bearings. Couplings. Back plate.

How are cooling tower fan blades attached? A. B. C. D.

9

Dial indicator Dial caliper Taper gauge Vernier dial indicator

Brazed Welded With U-bolts Coupling

When a high-speed gear mesh coupling is removed from a compressor, the coupling is typically match marked. What is the purpose of this? A. Ensures that it can be reassembled correctly. B. Ensures that it is disassembled correctly. C. Ensures that one coupling half is aligned to mesh.

10 Some centrifugal compressors use a high-speed gear-type coupling. How is this type of gear attached to the compressor shaft? A. B. C. D.

Bolted Welded Keyed Interference fit

11 With high-speed gear couplings, gear teeth showing a large polished area indicate which of the following? A. B. C. D.

Proper gear mesh Unit operation beyond recommended capacity Gear wear Galling


12 High-speed gear-type coupling for motor-to-centrifugal compressor connections are typically lubricated by: A. B. C. D.

Grease. Forced air. Process fluid. Oil.

13 Hose connections that are frequently made and “broken” would use which of the following types of couplings? A. B. C. D.

Flange Push-on Quick-connect Reusable

14 KAMAG generators may use any number of drivers. Which of the following coupling types are NOT recommended to tie the two together? A. B. C. D.

Flexible Gear Solid Both B and C

15 A common return signal wire shared by two or more signal wires is known as: A. B. C. D.

Impedance coupling. Magnetic coupling. Electrostatic coupling. Dielectric coupling.

16 Which of the following comes in two pieces? A. B. C. D.

Coupling Union Bushing Reducer

17 What type of fitting is used to connect tubing through a control panel or junction box? A. B. C. D.

Coupling Bulkhead Compression Cross


Benchmark your mechanical knowledge: Go to Appendix A for the correct answers and grade as follows: Safety: total right/20 = Lubrication: total right/20 = Bearings: total right/20 = Chain Drives: total right/20 = Belt Drives: total right/17 = Hydraulics: total right/16 = Couplings: total right/17 = Final Score: total of the scores/130 = 69% or below = Needs extensive mechanical training 70% –79% = Needs moderate mechanical training 80% –95% = Needs to review mechanical subject areas 95% –100% = Needs no training

4


Introduction As an industrial mechanic, you should always make safety your primary concern in the workplace. No job should be performed that would endanger your life or the lives of others. Always remember, “Safety first, safety last, safety always.” Each year in the United States more people are injured or killed in accidents at home, at work, school, at play, or while traveling than were injured or killed in either the Korean or Vietnam wars. Efforts to lessen or to eliminate the conditions that cause accidents are known as safety measures. Safety is a growing concern around the world, and safety skills are being taken more seriously today than ever before. People have come to realize that safety can be learned. Safety experts agree that it is possible to predict and prevent the majority of accidents. Few accidents simply “happen.” Most accidents are caused by ignorance, carelessness, neglect, or lack of skill. Since the introduction of automatic devices to move and handle materials, the exposure of workers to mechanical and handling hazards has been greatly reduced; however, the basic principles of safeguarding continue to be of great importance. Injuries result when loose clothing or hair is caught in rotating mechanisms; when fingers and hands are crushed in rollers, meshing gear teeth, belts, and chain drives; and when moving parts supply cutting, shearing, or crushing forces. Machine guards are designed to prevent such injuries. They may be fixed guards or automatic interlocking guards that prevent the operation of a machine unless a guard is in position at the danger point. Other types of guards prevent the operator from coming in contact with the dangerous part through a barrier, through devices that push the hands away, or through the use of sensor devices that stop the machine when hands are put into the danger zone. Making sure that the machinery, tools, and furniture associated with a job fit the workers who do that job is known as ergonomics, or human engineering. A properly designed workplace can reduce worker fatigue and increase safety on the job.

Safety First, Safety Always 51

Although many aspects of materials handling have been taken over by machine, a great deal of manual lifting and carrying must still be done in industry. In such cases, it is important to avoid unsafe work practices, such as improper lifting, carrying too heavy a load, or incorrect gripping. Workers must take responsibility for their own safety.

The Risk: Performing a Risk Assessment (the Preventive Management Tool) The identification and the analysis of risk, and making rational decisions based on the known risk, are the best preventive management tools a maintenance workforce can use. The risk assessment will allow a company to maximize its safety program and thus reduce accidents. The risk assessment process should be formalized in order to reduce accidents that could cause injuries, death, machine damage, and longer equipment stoppage. One must learn in maintenance that one must reduce risk at all times. In any task there are risk and thus the reduction or elimination of risk will make all tasks more successful and safe. In order to perform a risk assessment one must know the risk and then identify countermeasures to reduce or eliminate risk. Once one has reduced or eliminated the risk, a maintenance person can then perform the task with known acceptable risk. The Risk Assessment Worksheet shown below will assist in the explanation of this process.

Risk Assessment Worksheet The Risk Assessment Worksheet should be used for all maintenance tasks. What is known in maintenance is that a maintenance organization, in order to be successful, must have defined processes, and thus this worksheet is one of the important processes one must utilize in order to be successful. What makes this process so important is that lives are at risk if we do not follow it. Note: The score noted on the far right has three numbers to score each risk after countermeasures are in place. Once the assessment is completed then all scores are then added and matched to the Task Risk Scale at the bottom of the chart.

52 Safety First, Safety Always

WARNING: Sample Only—Table 4.1: This table is designed to provide a maintenance person with the understanding of the risk assessment process and should not be used for any determination of risk or safety criteria. Table 4.1 Risk Assessment Worksheet Risk Assessment Worksheet Task: Replace 100hp Electric Motor Date: 10/19/02 Name of Person Performing Risk Assessment: Ricky Smith

Risk

Potential safety hazard

Countermeasures

Score

Lifting Chain Break—single chain lift

Death, injury, equipment damage, loss of equipment run time

Inspect chain and safety latches to insure they do not have damage (replace if damaged; do not use) and are marked and identified to lift the direct lifting load you are about to lift (note lifting chains must have identified lifting capabilities marked by the chain manufacturer).

1

Lifting shackle does not have center pin and thus a grade 8 bolt will be used


None needed.

3

One maintenance person and one unskilled person changing the motor


Identify another person skilled and trained in rigging and hoisting.

1

Boom truck (crane truck)—Has not been load tested within the last 12 months


None needed.

3

Score total = 8


Risk = High/Danger Note: Task risk scale: The score for a risk is only determined once the countermeasures are followed; 1 = no risk; 2 = moderate risk; 3 = high risk (other countermeasures must be found). ANY 3 must be addressed immediately before the rigging and hoisting can begin. Total task risk: add all scores from each risk and match to the final score scale. Final score scale = minimal risk: 4–5; moderate risk (review with higher management before task begins): 6–7; high risk/danger (task will not begin until reduction of risk to level acceptable by management): 8 and above.

Risk Assessment Conclusion As one can establish from the chart above, if one can identify the known risk in a specific task then an organization can reduce or eliminate accidents. One can use the risk assessment process for not only large, complex tasks, but for all tasks. A simple laminated risk assessment card can be use by maintenance supervisors and maintenance personnel for emergency and just-in-time tasks. A maintenance planner should use the risk assessment process for all jobs planned. A copy of the Risk Assessment Worksheet should be a part of the job plan package given to maintenance technicians.

Lockout/Tagout/Tryout All machinery or equipment capable of movement must be de-energized or disengaged and blocked or locked out during cleaning, servicing, adjusting, or setting up operations, whenever required. The locking out of the control circuits in lieu of locking out main power disconnects is prohibited. All equipment control valve handles must be provided with a means for locking out. The lockout procedure requires that stored energy (i.e., mechanical, hydraulic, air) be released or blocked before equipment is locked out for repairs. Appropriate employees are provided with individually keyed personal safety locks. Employees are required to keep personal control of their key(s) while


they have safety locks in use. Employees must check the safety of the lockout by attempting a startup after making sure no one is exposed. Where the power disconnect does not also disconnect the electrical control circuit, the appropriate electrical enclosures must be identified. The control circuit can also be disconnected and locked out.

Manual Lifting Rules Manual lifting and handling of material must be done by methods that ensure the safety of both the employee and the material. It is important that employees whose work assignments require heavy lifting be trained and physically qualified by medical examination if necessary. The following are rules for manual lifting: Inspect for sharp edges, slivers, and wet or greasy spots. Wear gloves when lifting or handling objects with sharp or splintered edges. These gloves must be free of oil, grease, or other agents that may cause a poor grip. Inspect the route over which the load is to be carried. It should be free of obstructions or spillage that could cause tripping or slipping. Consider the distance the load is to be carried. Recognize the fact that your gripping power may weaken over long distances. Size up the load and make a preliminary “lift” to be sure the load is easily within your lifting capacity. If not, get help. If team lifting is required, personnel should be similar in size and physique. One person should act as leader and give the commands to lift, lower, etc. Two persons carrying a long piece of pipe or lumber should carry it on the same shoulder and walk in step. Shoulder pads may be used to prevent cutting shoulders and help reduce fatigue. To lift an object off the ground, the following are manual lifting steps: Make sure of good footing and set your feet about 10 to 15 inches apart. It may help to set one foot forward. Assume a knee-bend or squatting position, keeping your back straight and upright. Get a firm grip and lift the object by straightening your knees—not your back. Carry the load close to your body (not on extended arms). To turn or change your position, shift your feet—don’t twist your back. The steps for setting an object on the ground are the same as above, but reversed.


Power-Actuated Tools Employees using power-actuated tools must be properly trained. All poweractuated tools must be left disconnected until they are actually ready to be used. Each day before using, each power-actuated tool must be inspected for obstructions or defects. The power-actuated tool operators must have and use appropriate personal protective equipment such as hard hats, safety goggles, safety shoes, and ear protectors whenever they are using the equipment.

Machine Guarding Before operating any machine, the employee should have completed a training program on safe methods of operation. All machinery and equipment must be kept clean and properly maintained. There must be sufficient clearance provided around and between machines to allow for safe operations, setup, servicing, material handling, and waste removal. All equipment and machinery should be securely placed, and anchored when necessary, to prevent tipping or other movement that could result in damage or personal injury. Most machinery should be bolted to the floor to prevent falling during an earthquake. Also, the electrical cord should be fixed to a breaker or other shutoff device to stop power in case of machine movement. There should be a power shutoff switch within reach of the operator’s position. Electrical power to each machine must be capable of being locked out for maintenance, repair, or security. The noncurrent-carrying metal parts of electrically operated machines must be bonded and grounded. Foot-operated switches should be guarded and/or arranged to prevent accidental actuation by personnel or falling objects. All manually operated valves and switches controlling the operation of equipment and machines must be clearly identified and readily accessible. All EMERGENCY stop buttons should be colored RED. All the sheaves and


belts that are within 7 feet of the floor or working level should be properly guarded. All moving chains and gears must be properly guarded. All splash guards mounted on machines that use coolant must be positioned to prevent coolant from splashing the employees. The machinery guards must be secure and arranged so they do not present a hazard. All special hand tools used for placing and removing material must protect the operator’s hands. All revolving drums, barrels, and containers should be guarded by an enclosure that is interlocked with the drive mechanisms, so that revolution cannot occur unless the guard enclosure is in place. All arbors and mandrels must have firm and secure bearings and be free of play. A protective mechanism should be installed to prevent machines from automatically starting when power is restored after a power failure or shutdown. Machines should be constructed so as to be free from excessive vibration when under full load or mounted and running at full speed. If the machinery is cleaned with compressed air, the air must be pressure controlled, and personal protective equipment or other safeguards must be used to protect operators and other workers from eye and bodily injury. All fan blades should be protected by a guard having openings no larger than 12 inch when operating within 7 feet of the floor. Saws used for ripping equipment must be installed with antikickback devices and spreaders. All radial arm saws must be arranged so that the cutting head will gently return to the back of the table when released.

5

Rotor Balancing

Mechanical imbalance is one of the most common causes of machinery vibration and is present to some degree on nearly all machines that have rotating parts or rotors. Static, or standing, imbalance is the condition when there is more weight on one side of a centerline than the other. However, a rotor may be in perfect static balance and not be in a balanced state when rotating at high speed. If the rotor is a thin disk, careful static balancing may be accurate enough for high speeds. However, if the rotating part is long in proportion to its diameter, and the unbalanced portions are at opposite ends or in different planes, the balancing must counteract the centrifugal force of these heavy parts when they are rotating rapidly. This section provides information needed to understand and solve the majority of balancing problems using a vibration/balance analyzer, a portable device that detects the level of imbalance, misalignment, etc., in a rotating part based on the measurement of vibration signals.

Sources of Vibration due to Mechanical Imbalance Two major sources of vibration due to mechanical imbalance in equipment with rotating parts or rotors are: (1) assembly errors and (2) incorrect key length guesses during balancing.

Assembly Errors Even when parts are precision balanced to extremely close tolerances, vibration due to mechanical imbalance can be much greater than necessary due to assembly errors. Potential errors include relative placement of each part’s center of rotation, location of the shaft relative to the bore, and cocked rotors.

58 Rotor Balancing

Center of Rotation Assembly errors are not simply the additive effects of tolerances, but also include the relative placement of each part’s center of rotation. For example, a “perfectly” balanced blower rotor can be assembled to a “perfectly” balanced shaft, and yet the resultant imbalance can be high. This can happen if the rotor is balanced on a balancing shaft that fits the rotor bore within 0.5 mils (0.5 thousandths of an inch) and then is mounted on a standard cold-rolled steel shaft allowing a clearance of over 2 mils. Shifting any rotor from the rotational center on which it was balanced to the piece of machinery on which it is intended to operate can cause an assembly imbalance four to five times greater than that resulting simply from tolerances. For this reason, all rotors should be balanced on a shaft having a diameter as nearly the same as the shaft on which they will be assembled. For best results, balance the rotor on its own shaft rather than on a balancing shaft. This may require some rotors to be balanced in an overhung position, a procedure the balancing shop often wishes to avoid. However, it is better to use this technique rather than being forced to make too many balancing shafts. The extra precision balance attained by using this procedure is well worth the effort.

Method of Locating Position of Shaft Relative to Bore Imbalance often results with rotors that do not incorporate setscrews to locate the shaft relative to the bore (e.g., rotors that are end-clamped). In this case, the balancing shaft is usually horizontal. When the operator slides the rotor on the shaft, gravity causes the rotor’s bore to make contact at the 12 o’clock position on the top surface of the shaft. In this position, the rotor is end-clamped in place and then balanced. If the operator removes the rotor from the balancing shaft without marking the point of bore and shaft contact, it may not be in the same position when reassembled. This often shifts the rotor by several mils as compared to the axis on which it was balanced, thus causing an imbalance to be introduced. The vibration that results is usually enough to spoil what should have been a precision balance and produce a barely acceptable vibration level. In addition, if the resultant vibration is resonant with some part of the machine or structure, a more serious vibration could result. To prevent this type of error, the balancer operators and those who do final assembly should follow the following procedure. The balancer operator should permanently mark the location of the contact point between the

Rotor Balancing 59

bore and the shaft during balancing. When the equipment is reassembled in the plant or the shop, the assembler should also use this mark. For endclamped rotors, the assembler should slide the bore on the horizontal shaft, rotating both until the mark is at the 12 o’clock position, and then clamp it in place.

Cocked Rotor If a rotor is cocked on a shaft in a position different from the one in which it was originally balanced, an imbalanced assembly will result. If, for example, a pulley has a wide face that requires more than one setscrew, it could be mounted on-center, but be cocked in a different position than during balancing. This can happen by reversing the order in which the setscrews are tightened against a straight key during final mounting as compared to the order in which the setscrews were tightened on the balancing arbor. This can introduce a pure couple imbalance, which adds to the small couple imbalance already existing in the rotor and causes unnecessary vibration. For very narrow rotors (i.e., disk-shaped pump impellers or pulleys), the distance between the centrifugal forces of each half may be very small. Nevertheless, a very high centrifugal force, which is mostly counterbalanced statically by its counterpart in the other half of the rotor, can result. If the rotor is slightly cocked, the small axial distance between the two very large centrifugal forces causes an appreciable couple imbalance, which is often several times the allowable tolerance. This is due to the fact that the centrifugal force is proportional to half the rotor weight (at any one time, half of the rotor is pulling against the other half ) times the radial distance from the axis of rotation to the center of gravity of that half. To prevent this, the assembler should tighten each setscrew gradually—first one, then the other, and back again—so that the rotor is aligned evenly. On flange-mounted rotors such as flywheels, it is important to clean the mating surfaces and the bolt holes. Clean bolt holes are important because high couple imbalance can result from the assembly bolt pushing a small amount of dirt between the surfaces, cocking the rotor. Burrs on bolt holes also can produce the same problem.

Other There are other assembly errors that can cause vibration. Variances in bolt weights when one bolt is replaced by one of a different length or material

60 Rotor Balancing

can cause vibration. For setscrews that are 90 degrees apart, the tightening sequence may not be the same at final assembly as during balancing. To prevent this, the balancer operator should mark which was tightened first.

Key Length With a keyed-shaft rotor, the balancing process can introduce machine vibration if the assumed key length is different from the length of the one used during operation. Such an imbalance usually results in a mediocre or “good” running machine as opposed to a very smooth running machine. For example, a “good” vibration level that can be obtained without following the precautions described in this section is amplitude of 0.12 inches/second (3.0 mm/sec.). By following the precautions, the orbit can be reduced to about 0.04 in./sec. (1 mm/sec.). This smaller orbit results in longer bearing or seal life, which is worth the effort required to make sure that the proper key length is used. When balancing a keyed-shaft rotor, one half of the key’s weight is assumed to be part of the shaft’s male portion. The other half is considered to be part of the female portion that is coupled to it. However, when the two rotor parts are sent to a balancing shop for rebalancing, the actual key is rarely included. As a result, the balance operator usually guesses at the key’s length, makes up a half key, and then balances the part. (Note: A “half key” is of full-key length, but only half-key depth.) In order to prevent an imbalance from occurring, do not allow the balance operator to guess the key length. It is strongly suggested that the actual key length be recorded on a tag that is attached to the rotor to be balanced. The tag should be attached in such a way that another device (such as a coupling half, pulley, fan, etc.) cannot be attached until the balance operator removes the tag.

Theory of Imbalance Imbalance is the condition in which there is more weight on one side of a centerline than the other. This condition results in unnecessary vibration, which generally can be corrected by the addition of counterweights. There are four types of imbalance: (1) static, (2) dynamic, (3) coupled, and (4) dynamic imbalance combinations of static and couple.

Rotor Balancing 61

Static Static imbalance is single-plane imbalance acting through the center of gravity of the rotor, perpendicular to the shaft axis. The imbalance also can be separated into two separate single-plane imbalances, each acting in-phase or at the same angular relationship to each other (i.e., 0 degrees apart). However, the net effect is as if one force is acting through the center of gravity. For a uniform straight cylinder such as a simple paper machine roll or a multigrooved sheave, the forces of static imbalance measured at each end of the rotor are equal in magnitude (i.e., the ounce-inches or gramcentimeters in one plane are equal to the ounce-inches or gram-centimeters in the other). In static imbalance, the only force involved is weight. For example, assume that a rotor is perfectly balanced and, therefore, will not vibrate regardless of the speed of rotation. Also assume that this rotor is placed on frictionless rollers or “knife edges.” If a weight is applied on the rim at the center of gravity line between two ends, the weighted portion immediately rolls to the 6 o’clock position due to the gravitational force. When rotation occurs, static imbalance translates into a centrifugal force. As a result, this type of imbalance is sometimes referred to as “force imbalance,” and some balancing machine manufacturers use the word “force” instead of “static” on their machines. However, when the term “force imbalance” was just starting to be accepted as the proper term, an American standardization committee on balancing terminology standardized the term “static” instead of “force.” The rationale was that the role of the standardization committee was not to determine and/or correct right or wrong practices, but to standardize those currently in use by industry. As a result, the term “static imbalance” is now widely accepted as the international standard and, therefore, is the term used here.

Dynamic Dynamic imbalance is any imbalance resolved to at least two correction planes (i.e., planes in which a balancing correction is made by adding or removing weight). The imbalance in each of these two planes may be the result of many imbalances in many planes, but the final effects can be limited to only two planes in almost all situations. An example of a case where more than two planes are required is flexible rotors (i.e., long rotors running at high speeds). High speeds are considered

62 Rotor Balancing

to be revolutions per minute (rpm) higher than about 80% of the rotor’s first critical speed. However, in over 95% of all run-of-the-mill rotors (e.g., pump impellers, armatures, generators, fans, couplings, pulleys, etc.), two-plane dynamic balance is sufficient. Therefore, flexible rotors are not covered in this document because of the low number in operation and the fact that specially trained people at the manufacturer’s plant almost always perform balancing operations. In dynamic imbalance, the two imbalances do not have to be equal in magnitude to each other, nor do they have to have any particular angular reference to each other. For example, they could be 0 (in-phase), 10, 80, or 180 degrees from each other. Although the definition of dynamic imbalance covers all two-plane situations, an understanding of the components of dynamic imbalance is needed so that its causes can be understood. Also, an understanding of the components makes it easier to understand why certain types of balancing do not always work with many older balancing machines for overhung rotors and very narrow rotors. The primary components of dynamic imbalance include: number of points of imbalance, amount of imbalance, phase relationships, and rotor speed.

Points of Imbalance The first consideration of dynamic balancing is the number of imbalance points on the rotor, as there can be more than one point of imbalance within a rotor assembly. This is especially true in rotor assemblies with more than one rotating element, such as a three-rotor fan or multistage pump.

Amount of Imbalance The amplitude of each point of imbalance must be known to resolve dynamic balance problems. Most dynamic balancing machines or in situ balancing instruments are able to isolate and define the specific amount of imbalance at each point on the rotor.

Phase Relationship The phase relationship of each point of imbalance is the third factor that must be known. Balancing instruments isolate each point of imbalance and determine their phase relationship. Plotting each point of imbalance on a polar plot does this. In simple terms, a polar plot is a circular display of the

Rotor Balancing 63

shaft end. Each point of imbalance is located on the polar plot as a specific radial, ranging from 0 to 360 degrees.

Rotor Speed Rotor speed is the final factor that must be considered. Most rotating elements are balanced at their normal running speed or over their normal speed range. As a result, they may be out of balance at some speeds that are not included in the balancing solution. As an example, the wheel and tires on your car are dynamically balanced for speeds ranging from zero to the maximum expected speed (i.e., eighty miles per hour). At speeds above eighty miles per hour, they may be out of balance.

Coupled Coupled imbalance is caused by two equal noncollinear imbalance forces that oppose each other angularly (i.e., 180 degrees apart). Assume that a rotor with pure coupled imbalance is placed on frictionless rollers. Because the imbalance weights or forces are 180 degrees apart and equal, the rotor is statically balanced. However, a pure coupled imbalance occurs if this same rotor is revolved at an appreciable speed. Each weight causes a centrifugal force, which results in a rocking motion or rotor wobble. This condition can be simulated by placing a pencil on a table, then at one end pushing the side of the pencil with one finger. At the same time, push in the opposite direction at the other end. The pencil will tend to rotate end-over-end. This end-over-end action causes two imbalance “orbits,” both 180 degrees out of phase, resulting in a “wobble” motion.

Dynamic Imbalance Combinations of Static and Coupled Visualize a rotor that has only one imbalance in a single plane. Also visualize that the plane is not at the rotor’s center of gravity, but is off to one side. Although there is no other source of couple, this force to one side of the rotor not only causes translation (parallel motion due to pure static imbalance), but also causes the rotor to rotate or wobble end-over-end as from a couple. In other words, such a force would create a combination of both static and couple imbalance. This again is dynamic imbalance. In addition, a rotor may have two imbalance forces exactly 180 degrees opposite to each other. However, if the forces are not equal in magnitude,

64 Rotor Balancing

the rotor has a static imbalance in combination with its pure couple. This combination is also dynamic imbalance. Another way of looking at it is to visualize the usual rendition of dynamic imbalance—imbalance in two separate planes at an angle and magnitude relative to each other not necessarily that of pure static or pure couple. For example, assume that the angular relationship is 80 degrees and the magnitudes are 8 units in one plane and 3 units in the other. Normally, you would simply balance this rotor on an ordinary two-plane dynamic balancer and that would be satisfactory. But for further understanding of balancing, imagine that this same rotor is placed on static balancing rollers, whereby gravity brings the static imbalance components of this dynamically out-of-balance rotor to the 6 o’clock position. The static imbalance can be removed by adding counter-balancing weights at the 12 o’clock position. Although statically balanced, however, the two remaining forces result in a pure coupled imbalance. With the entire static imbalance removed, these two forces are equal in magnitude and exactly 180 degrees apart. The coupled imbalance can be removed, as with any other coupled imbalance, by using a two-plane dynamic balancer and adding counterweights. Note that whenever you hear the word “imbalance,” you should mentally add the word “dynamic” to it. Then when you hear “dynamic imbalance,” mentally visualize “combination of static and coupled imbalance.” This will be of much help not only in balancing, but in understanding phase and coupling misalignment as well.

Balancing Imbalance is one of the most common sources of major vibration in machinery. It is the main source in about 40% of the excessive vibration situations. The vibration frequency of imbalance is equal to one times the rpm (l × rpm) of the imbalanced rotating part. Before a part can be balanced using the vibration analyzer, certain conditions must be met: ●

The vibration must be due to mechanical imbalance;

●

Weight corrections can be made on the rotating component.

Rotor Balancing 65

In order to calculate imbalance units, simply multiply the amount of imbalance by the radius at which it is acting. In other words, one ounce of imbalance at a one-inch radius will result in one oz.-in. of imbalance. Five ounces at one-half inch radius results in 2 21 oz.-in. of imbalance. (Dynamic imbalance units are measured in ounce-inches [oz.-in.] or gram-millimeters [g.-mm.].) Although this refers to a single plane, dynamic balancing is performed in at least two separate planes. Therefore, the tolerance is usually given in single-plane units for each plane of correction. Important balancing techniques and concepts to be discussed in the sections to follow include: in-place balancing, single-plane versus two-plane balancing, precision balancing, techniques that make use of a phase shift, and balancing standards.

In-Place Balancing In most cases, weight corrections can be made with the rotor mounted in its normal housing. The process of balancing a part without taking it out of the machine is called in-place balancing. This technique eliminates costly and time consuming disassembly. It also prevents the possibility of damage to the rotor, which can occur during removal, transportation to and from the balancing machine, and reinstallation in the machine.

Single-Plane versus Two-Plane Balancing The most common rule of thumb is that a disk-shaped rotating part usually can be balanced in one correction plane only, whereas parts that have appreciable width require two-plane balancing. Precision tolerances, which become more meaningful for higher performance (even on relatively narrow face width), suggest two-plane balancing. However, the width should be the guide, not the diameter-to-width ratio. For example, a 20" wide rotor could have a large enough couple imbalance component in its dynamic imbalance to require two-plane balancing. (Note: The couple component makes two-plane balancing important.) Yet, if the 20" width is on a rotor of large diameter that qualifies as a “disk-shaped rotor,” even some of the balance manufacturers erroneously would call for a single-plane balance. It is true that the narrower the rotor, the less the chance for a large couple component and, therefore, the greater the possibility of getting by with a single-plane balance. For rotors over 4" to 5" in width, it is best to check

66 Rotor Balancing

for real dynamic imbalance (or for couple imbalance). Unfortunately, you cannot always get by with a static- and couple-type balance, even for very narrow flywheels used in automobiles. Although most of the flywheels are only 1" to 1 12 " wide, more than half have enough couple imbalance to cause excessive vibration. This obviously is not due to a large distance between the planes (width), but due to the fact that the flywheel’s mounting surface can cause it to be slightly cocked or tilted. Instead of the flywheel being 90 degrees to the shaft axis, it may be perhaps 85 to 95 degrees, causing a large couple despite its narrow width. This situation is very common with narrow and disc-shaped industrial rotors such as single-stage turbine wheels, narrow fans, and pump impellers. The original manufacturer often accepts the guidelines supplied by others and performs a single-plane balance only. By obtaining separate readings for static and couple, the manufacturer could and should easily remove the remaining couple. An important point to remember is that static imbalance is always removed first. In static and couple balancing, remove the static imbalance first, and then remove the couple.

Precision Balancing Most original-equipment manufacturers balance to commercial tolerances, a practice that has become acceptable to most buyers. However, due to frequent customer demands, some of the equipment manufacturers now provide precision balancing. Part of the driving force for providing this service is that many large mills and refineries have started doing their own precision balancing to tolerances considerably closer than those used by the original-equipment manufacturer. For example, the International Standards Organization (ISO) for process plant machinery calls for a G6.3 level of balancing in its balancing guide. This was calculated based on a rotor running free in space with a restraint vibration of 6.3 mm/sec. (0.25 in./sec.) vibration velocity. Precision balancing requires a G2.5 guide number, which is based on 2.5 mm/sec. (0.1 in./sec.) vibration velocity. As can be seen from this, 6.3 mm/sec. (0.25 in./sec.) balanced rotors will vibrate more than the 2.5 mm/sec. (0.1 in./sec.) precision balanced rotors. Many vibration guidelines now consider 2.5 mm/sec. (0.1 in./sec.) “good,” creating the demand for precision balancing. Precision balancing tolerances can produce velocities of 0.01 in./sec. (0.3 mm/sec.) and lower.

Rotor Balancing 67

It is true that the extra weight of nonrotating parts (i.e., frame and foundation) reduces the vibration somewhat from the free-in-space amplitude. However, it is possible to reach precision balancing levels in only two or three additional runs, providing the smoothest running rotor. The extra effort to the balance operator is minimal because he already has the “feel” of the rotor and has the proper setup and tools in hand. In addition, there is a large financial payoff for this minimal extra effort due to decreased bearing and seal wear.

Techniques Using Phase Shift If we assume that there is no other source of vibration other than imbalance (i.e., we have perfect alignment, a perfectly straight shaft, etc.), it is readily seen that pure static imbalance gives in-phase vibrations, and pure coupled imbalance gives various phase relationships. Compare the vertical reading of a bearing at one end of the rotor with the vertical reading at the other end of the rotor to determine how that part is shaking vertically. Then compare the horizontal reading at one end with the horizontal reading at the other end to determine how the part is shaking horizontally. If there is no resonant condition to modify the resultant vibration phase, then the phase for both vertical and horizontal readings is essentially the same even though the vertical and horizontal amplitudes do not necessarily correspond. In actual practice, this may be slightly off due to other vibration sources such as misalignment. In performing the analysis, what counts is that when the source of the vibration is primarily from imbalance, then the vertical reading phase differences between one end of the rotor and the other will be very similar to the phase differences when measured horizontally. For example, vibrations 60 degrees out of phase vertically would show 60 degrees out of phase horizontally within 20%. However, the horizontal reading on one bearing will not show the same phase relationship as the vertical reading on the same bearing. This is due to the pickup axis being oriented in a different angular position, as well as the phase adjustment due to possible resonance. For example, the horizontal vibration frequency may be below the horizontal resonance of various major portions of machinery, whereas the vertical vibration frequency may be above the natural frequency of the floor supporting the machine. First, determine how the rotor is vibrating vertically by comparing “vertical only” readings with each other. Then, determine how the rotor is vibrating horizontally. If, the rotor is shaking horizontally and vertically and the phase

68 Rotor Balancing

differences are relatively similar, then the source of vibration is likely to be imbalance. However, before coming to a final conclusion, be sure that other l × rpm sources (e.g., bent shaft, eccentric armature, misaligned coupling) are not at fault.

Balancing Standards The ISO has published standards for acceptable limits for residual imbalance in various classifications of rotor assemblies. Balancing standards are given in oz-in. or lb-in. per pound of rotor weight or the equivalent in metric units (g-mm/kg). The oz-in. are for each correction plane for which the imbalance is measured and corrected. Caution must be exercised when using balancing standards. The recommended levels are for residual imbalance, which is defined as imbalance of any kind that remains after balancing.

Balancing of Rotating Machinery 100,000

G8

30

1

G2

50 G1 00

0.1

10,000

G4

1,000

0

0.01

G1 6

100

G6

.3

0.001

G2

.5 G1

0.0001

10

G0 4

1

0.000010 100

1000 Speed, RPM

10,000

0.1

Acceptable Residual Unbalance per Unit of Rotor Weight, gm mm/kg

Acceptable Residual Unbalance per Unit of Rotor Weight, LB-IN./LB

Figure 5.1 and Table 5.1 are the norms established for most rotating equipment. Additional information can be obtained from ISO 5406 and 5343.

Figure 5.1 Balancing standards: residual imbalance per unit rotor weight

Rotor Balancing 69

Table 5.1 Balance quality grades for various groups of rigid rotors Balance quality grade G4,000 G1,600 G630 G250 G100

G40

G16 G6.3

G2.5

G1 G0.4

Type of rotor Crankshaft drives of rigidly mounted slow marine diesel engines with uneven number of cylinders. Crankshaft drives of rigidly mounted large two-cycle engines. Crankshaft drives of rigidly mounted large four-cycle engines; crankshaft drives of elastically mounted marine diesel engines. Crankshaft drives of rigidly mounted fast four-cylinder diesel engines. Crankshaft drives of fast diesel engines with six or more cylinders; complete engines (gasoline or diesel) for cars and trucks. Car wheels, wheel rims, wheel sets, drive shafts; crankshaft drives of elastically mounted fast four-cycle engines (gasoline and diesel) with six or more cylinders; crankshaft drives for engines of cars and trucks. Parts of agricultural machinery; individual components of engines (gasoline or diesel) for cars and trucks. Parts or process plant machines; marine main-turbine gears; centrifuge drums; fans; assembled aircraft gas-turbine rotors; flywheels; pump impellers; machine-tool and general machinery parts; electrical armatures. Gas and steam turbines; rigid turbo-generator rotors; rotors; turbo-compressors; machine-tool drives; small electrical armatures; turbine-driven pumps. Tape recorder and phonograph drives; grinding-machine drives. Spindles, disks, and armatures of precision grinders; gyroscopes.

Similar standards are available from the American National Standards Institute (ANSI) in their publication ANSI S2.43-1984. So far, there has been no consideration of the angular positions of the usual two points of imbalance relative to each other or the distance between the two correction planes. For example, if the residual imbalances in each of the two planes were in-phase, they would add to each other to create more static imbalance.

70 Rotor Balancing

Most balancing standards are based on a residual imbalance and do not include multiplane imbalance. If they are approximately 180 degrees to each other, they form a couple. If the distance between the planes is small, the resulting couple is small; if the distance is large, the couple is large. A couple creates considerably more vibration than when the two residual imbalances are in-phase. Unfortunately, there is nothing in the balancing standards that takes this into consideration. There is another problem that could also result in excessive imbalancerelated vibration even though the ISO standards were met. The ISO standards call for a balancing grade of G6.3 for components such as pump impellers, normal electric armatures, and parts of process plant machines. This results in an operating speed vibration velocity of 6.3 mm/sec. (0.25 in./sec.) vibration velocity. However, practice has shown that an acceptable vibration velocity is 0.1 in./sec. and the ISO standard of G2.5 is really required. As a result of these discrepancies, changes in the recommended balancing grade are expected in the future.

6

Bearings

A bearing is a machine element that supports a part, such as a shaft, that rotates, slides, or oscillates in or on it. There are two broad classifications of bearings: plain and rolling element (also called antifriction). Plain bearings are based on sliding motion made possible through the use of a lubricant. Antifriction bearings are based on rolling motion, which is made possible by balls or other types of rollers. In modern rotor systems operating at relatively high speeds and loads, the proper selection and design of the bearings and bearing-support structure are key factors affecting system life.

Types of Movement The type of bearing used in a particular application is determined by the nature of the relative movement and other application constraints. Movement can be grouped into the following categories: rotation about a point, rotation about a line, translation along a line, rotation in a plane, and translation in a plane. These movements can be either continuous or oscillating. Although many bearings perform more than one function, they can generally be classified based on types of movement. There are three major classifications of both plain and rolling element bearings: radial, thrust, and guide. Radial bearings support loads that act radially and at right angles to the shaft center line. These loads may be visualized as radiating into or away from a center point like the spokes on a bicycle wheel. Thrust bearings support or resist loads that act axially. These may be described as endwise loads that act parallel to the center line toward the ends of the shaft. This type of bearing prevents lengthwise or axial motion of a rotating shaft. Guide bearings support and align members having sliding or reciprocating motion. This type of bearing guides a machine element in its lengthwise motion, usually without rotation of the element. Table 6.1 gives examples of bearings that are suitable for continuous movement; Table 6.2 shows bearings that are appropriate for oscillatory movement only. For the bearings that allow movements in addition to the one listed, the effect on machine design is described in the column “Effect

Table 6.1 Bearing selection guide (continuous movement)

Constraint applied to the movement

Examples of arrangements which allow movement only within this constraint

Examples of arrangements which allow movement but also have other degrees of freedom

About a point

Gimbals

Ball on a recessed plate

Ball must be forced into contact with the plate

About a line

Journal bearing with double thrust location

Journal bearing

Simple journal bearing allows free axial movement as well

Double conical bearing

Screw and nut

Gives some related axial movement as well

Ball joint or spherical roller bearing

Allows some angular freedom to the line of rotation

Effect of the other degrees of freedom

Table 6.1 continued

About a line

Crane wheels restrained between two rails

Railway or crane wheel on a track

These arrangements need to be loaded into contact. This is usually done by gravity. Wheels on a single rail or cable need restraint to prevent rotation about the track member

Pulley wheel on a cable

Hovercraft or hoverpad on a track

In a plane (rotation)

In a plane (translation)

Double thrust bearing

Single thrust bearing

Single thrust bearing must be loaded into contact

Hovercraft or hoverpad

Needs to be loaded into contact usually by gravity

Source:M.J. Neale, Society of Automotive Engineers Inc. Bearings—A Tribology Handbook. Oxford: Butterworth–Heinemann, 1993.

Table 6.2 Bearing selection guide (oscillatory movement)

Constraint applied to the movement

Examples of arrangements which allow movement only within this constraint

Examples of arrangements which allow this movement but also have other degrees of freedom

Effect of the other degrees of freedom

About a point

Hookes joint

Cable connection between components

Cable needs to be kept in tension

About a line

Crossed strip flexure pivot

Torsion suspension

A single torsion suspension gives no lateral location

Knife-edge pivot

Must be loaded into contact

Rubber bush

Gives some axial and lateral flexibility as well

Rocker pad

Gives some related translation as well. Must be loaded into contact

Table 6.2 continued

Along a line

Crosshead and guide bars

In a plane (rotation)

In a plane (translation)

Plate between upper and lower guide blocks

Piston and cylinder

Piston can rotate as well unless it is located by connecting rod

Rubber ring or disc

Gives some axial and lateral flexibility as well

Block sliding on a plate

Must be loaded into contact

Source:M.J. Neale, Society of Automotive Engineers Inc. Bearings—A Tribology Handbook. Oxford: Butterworth–Heinemann, 1993.

76 Bearings

Table 6.3 Comparison of plain and rolling element bearings Rolling element

Plain

Assembly on crankshaft is virtually impossible, except with very short or built-up crankshafts Cost relatively high Hardness of shaft unimportant

Assembly on crankshaft is no problem as split bearings can be used

Heavier than plain bearings Housing requirement not critical Less rigid than plain bearings Life limited by material fatigue Lower friction results in lower power consumption Lubrication easy to accomplish; the required flow is low except at high speed

Noisy operation Poor tolerance of shaft deflection Poor tolerance of hard dirt particles

Requires more overall space: Length: Smaller than plain Diameter: Larger than plain Running Friction: Very low at low speeds May be high at high speeds Smaller radial clearance than plain

Source: Integrated Systems Inc.

Cost relatively low Hardness of shaft important with harder bearings Lighter than rolling element bearings Rigidity and clamping most important housing requirement More rigid than rolling element bearings Life not generally limited by material fatigue Higher friction causes more power consumption Lubrication pressure feed critically important; required flow is large, susceptible to damage by contaminants and interrupted lubricant flow Quiet operation Moderate tolerance of shaft deflection Moderate tolerance of dirt particles, depending on hardness of bearing Requires less overall space: Length: Larger than rolling element Diameter: Smaller than rolling element Running Friction: Higher at low speeds Moderate at usual crank speeds Larger radial clearance than rolling element

Bearings 77

of the other degrees of freedom.” Table 6.3 compares the characteristics, advantages, and disadvantages of plain and rolling element bearings.

About a Point (Rotational) Continuous movement about a point is rotation, a motion that requires repeated use of accurate surfaces. If the motion is oscillatory rather than continuous, some additional arrangements must be made in which the geometric layout prevents continuous rotation.

About a Line (Rotational) Continuous movement about a line is also referred to as rotation, and the same comments apply as for movement about a point.

Along a Line (Translational) Movement along a line is referred to as translation. One surface is generally long and continuous, and the moving component is usually supported on a fluid film or rolling contact in order to achieve an acceptable wear rate. If the translational movement is reciprocation, the application makes repeated use of accurate surfaces, and a variety of economical bearing mechanisms are available.

In a Plane (Rotational/Translational) If the movement in a plane is rotational or both rotational and oscillatory, the same comments apply as for movement about a point. If the movement in a plane is translational or both translational and oscillatory, the same comments apply as for movement along a line.

Commonly Used Bearing Types As mentioned before, the major bearing classifications are plain and rolling element. These types of bearings are discussed in the sections to follow. Table 6.4 is a bearings characteristics summary. Table 6.5 is a selection guide for bearings operating with continuous rotation and special environmental conditions. Table 6.6 is a selection guide for bearings operating with continuous rotation and special performance requirements. Table 6.7 is a selection

78 Bearings

Table 6.4 Bearings characteristic summary Bearing type

Description

Plain Lobed

See Table 6.3. See Radial, elliptical.

Radial or journal Cylindrical Elliptical Four-axial grooved Partial arc Tilting pad

Thrust

Rolling element Ball Single-row Radial nonfilling slot

Radial filling slot Angular contact radial thrust Ball-thrust Double-row Double-roll, self-aligning Roller

Gas lubricated, low-speed applications. Oil lubricated, gear and turbine applications, stiffer and somewhat more stable bearing. Oil lubricated, higher-speed applications than cylindrical. Not a bearing type, but a theoretical component of grooved and lobed bearing configurations. High-speed applications where hydrodynamic instability and misalignment are common problems. Semifluid lubrication state, relatively high friction, lower service pressures with multicollar version, used at low speeds. See Table 6.3. Radial and axial loads, moderate- to high-speed applications. Higher speed and lighter load applications than roller bearings. Also referred to as Conrad or deep-groove bearing. Sustains combined radial and thrust loads, or thrust loads alone, in either direction, even at high speeds. Not self-aligning. Handles heavier loads than nonfilling slot. Radial loads combined with thrust loads, or heavy thrust loads alone. Axial deflection must be limited. Very high thrust loads in one direction only, no radial loading, cannot be operated at high speeds. Heavy radial with minimal bearing deflection and light thrust loads. Moderate radial and limited thrust loads. Handles heavier loads and shock better than ball bearings, but are more limited in speed than ball bearings. Continued

Bearings 79

Table 6.4 continued Bearing type

Description

Cylindrical

Heavy radial loads, fairly high speeds, can allow free axial shaft movement. Does not normally support thrust loads, used in space-limited applications, angular mounting of rolls in double-row version tolerates combined axial and thrust loads. High radial and moderate-to-heavy thrust loads, usually comes in double-row mounting that is inherently self-aligning. Heavy radial and thrust loads. Can be preloaded for maximum system rigidity.

Needle-type cylindrical or barrel

Spherical

Tapered

Source: Integrated Systems, Inc.

guide for oscillating movement and special environment or performance requirements.

Plain Bearings All plain bearings also are referred to as fluid-film bearings. In addition, radial plain bearings also are commonly referred to as journal bearings. Plain bearings are available in a wide variety of types or styles and may be self-contained units or built into a machine assembly. Table 6.8 is a selection guide for radial and thrust plain bearings. Plain bearings are dependent on maintaining an adequate lubricant film to prevent the bearing and shaft surfaces from coming into contact, which is necessary to prevent premature bearing failure. However, this is difficult to achieve, and some contact usually occurs during operation. Material selection plays a critical role in the amount of friction and the resulting seizure and wear that occurs with surface contact. Note that fluid-film bearings do not have the ability to carry the full load of the rotor assembly at any speed and must have turning gear to support the rotor’s weight at low speeds.

Thrust or Fixed Thrust plain bearings consist of fixed shaft shoulders or collars that rest against flat bearing rings. The lubrication state may be semifluid, and friction

External vibration

Bearing type

High temp.

Low temp.

Vacuum

Wet/humid

Dirt/dust

Plain, externally pressurized

1 (With gas lubrication)

2

2

4 (Lubricant oxidizes)

3 (May have high starting torque)

2 (1 when gas lubricated) Seals essential

1

Plain, porous metal (oil impregnated) Plain, rubbing(nonmetallic) Plain, fluid film

2 (Up to temp. limit of material) 2 (Up to temp. limit of lubricant) Consult manufacturer above 150◦ C Effect of thermal expansion on fits

2

No (affected by lubricant feed) Possible with special lubricant 1

2 (Shaft must not corrode)

2 (Seals help)

2

Possible with special lubricant 3 (With special lubricant)

2

2 (With seals and filtration) Sealing essential

2

Rolling

Things to watch with all bearings

2 (May have high starting torque) 2

Effect of thermal expansion on fits

2

3 (With seals)

Corrosion

Rating: 1 - Excellent, 2 - Good, 3 - Fair, 4 - Poor Source: Adapted by Integrated Systems, Inc. from Bearings—A Tribology Handbook, M.J. Neale, Society of AutomotiveEngineers, Inc., Butterworth–Heinemann Ltd., Oxford, Great Britain, 1993.

2

3 (Consult manufacturers) Fretting

80 Bearings

Table 6.5 Bearings selection guide for special environmental conditions (continuous rotation)

Table 6.6 Bearing selection guide for particular performance requirements (continuous rotation)

Bearing type

Accurate radial location

Plain, externally pressurized

1

Plain, fluid film

3

Plain, porous metal (oil impregnated) Plain, rubbing (nonmetallic) Rolling

2

4 2

Axial load Low starting Silent capacity as well torque running

Standard parts available

Simple lubrication

No (need separate thrust bearing) No (need separate thrust bearing) Some

1

1

No

4 (Need special system)

2

1

Some

2 (Usually requires circulation system)

2

1

Yes

1

Some in most instances Yes in most instances

4

3

Some

1

1

Usually satis- Yes factory

2 (When grease lubricated)

Rating: 1 - Excellent, 2 - Good, 3 - Fair, 4 - Poor Bearings 81

Source: Adapted by Integrated Systems Inc. from M. J. Neale, Society of Automotive Engineers Inc. Bearings—A Tribology Handbook. Oxford: Butterworth–Heinemann, 1993.

82 Bearings

Table 6.7 Bearing selection guide for special environments or performance (oscillating movement) Bearing type

High temp.

Low temp.

Low friction

Wet/humid

Dirt/dust

External vibration

Knife edge pivots

2

2

1

2

4

Plain, porous metal (oil impregnated) Plain, rubbing

3 (Friction can be high) 1

2

2

2 (With PTFE)

2

1

Sealing essential 2 (Sealing helps) Sealing essential

Rubber bushes

4 (Lubricant oxidizes) 2 (Up to temp. limit of material) Consult manufacturer above 150◦ C 4

2 (Watch corrosion) 2

4

Strip flexures

2

1

Elastically stiff 1

Rolling

2 (Shaft must not corrode) 2 (With seals)

1 4

1

1

1

2 (Watch corrosion)

1

1

Rating: 1 - Excellent, 2 - Good, 3 - Fair, 4 - Poor Source: Adapted by Integrated Systems Inc. from M.J. Neale, Society of Automotive Engineers Inc. Bearings—A Tribology Handbook. Oxford: Butterworth–Heinemann, 1993.

Bearings 83

Table 6.8 Plain bearing selection guide Journal bearings Characteristics

Direct lined

Insert liners

Accuracy

Dependent upon facilities and skill available Doubtful

Precision components Consistent

Initial cost may be lower

Initial cost may be higher Easily done by replacement Ability to sustain higher peak loads Extensive range available

Quality (consistency) Cost Ease of Repair

Difficult and costly

Condition upon extensive use

Likely to be weak in fatigue

Materials used

Limited to white metals

Thrust bearings Characteristics

Flanged journal bearings

Cost

Costly to manufacture

Replacement

Involves whole journal/thrust component

Materials used

Thrust face materials limited in larger sizes Aids assembly on a production line

Benefits

Separate thrust washer Much lower initial cost Easily replaced without moving journal bearing Extensive range available Aligns itself with the housing

Source: Adapted by Integrated Systems Inc. from M.J. Neale, Society of Automotive Engineers Inc. Bearings—A Tribology Handbook. Oxford: Butterworth–Heinemann, 1993.

is relatively high. In multicollar thrust bearings, allowable service pressures are considerably lower because of the difficulty in distributing the load evenly between several collars. However, thrust ring performance can be improved by introducing tapered grooves. Figure 6.1 shows a mounting half-section for a vertical thrust bearing.

84 Bearings

Air vent

Thrust block Oil retainer

Oil level

Self aligning equalizing base

Runner Shoe

Oil tight joint

Radial bearing

Mach. frame

Figure 6.1 Half section of mounting for vertical thrust bearing

Radial or Journal Plain radial, or journal, bearings also are referred to as sleeve or Babbit bearings. The most common type is the full journal bearing, which has 360-degree contact with its mating journal. The partial journal bearing has less than 180-degree contact and is used when the load direction is constant. The sections to follow describe the major types of fluid-film journal bearings: plain cylindrical, four-axial groove, elliptical, partial-arc, and tilting-pad.

Plain Cylindrical The plain cylindrical journal bearing (Figure 6.2) is the simplest of all journal bearing types. The performance characteristics of cylindrical bearings are well established, and extensive design information is available. Practically, use of the unmodified cylindrical bearing is generally limited to gas-lubricated bearings and low-speed machinery. Four-Axial Groove Bearing To make the plain cylindrical bearing practical for oil or other liquid lubricants, it is necessary to modify it by the addition of grooves or holes through which the lubricant can be introduced. Sometimes, a single circumferential

Bearings 85

Load

Journal

Dia

me

ter

Bearing

D

Clearance C

Figure 6.2 Plain cylindrical bearing Load 35°

35°

Groove

Dia

Journal

me

ter

Bearing

D Clearance C

9° 45°

45°

Figure 6.3 Four-axial groove bearing groove in the middle of the bearing is used. In other cases, one or more axial grooves are provided. The four-axial groove bearing is the most commonly used oil-lubricated sleeve bearing. The oil is supplied at a nominal gauge pressure that ensures an adequate oil flow and some cooling capability. Figure 6.3 illustrates this type of bearing.

Elliptical Bearing The elliptical bearing is oil-lubricated and typically is used in gear and turbine applications. It is classified as a lobed bearing in contrast to a grooved bearing. Where the grooved bearing consists of a number of partial arcs with a common center, the lobed bearing is made up of partial arcs whose centers

86 Bearings

Load

Journal Radius R Bearing

Groove R

R+ C

C

R+

mC mC

Journal

15° 15°

C = Clearance m = Ellipticity

Figure 6.4 Elliptical bearing

do not coincide. The elliptical bearing consists of two partial arcs where the bottom arc has its center a distance above the bearing center. This arrangement has the effect of preloading the bearing, where the journal center eccentricity with respect to the loaded arc is increased and never becomes zero. This results in the bearing being stiffened, somewhat improving its stability. An elliptical bearing is shown in Figure 6.4.

Partial-Arc Bearings A partial-arc bearing is not a separate type of bearing. Instead, it refers to a variation of previously discussed bearings (e.g., grooved and lobed bearings) that incorporates partial arcs. It is necessary to use partial-arc bearing data to incorporate partial arcs in a variety of grooved and lobed bearing configurations. In all cases, the lubricant is a liquid and the bearing film is laminar. Figure 6.5 illustrates a typical partial-arc bearing.

Tilting-Pad Bearings Tilting-pad bearings are widely used in high-speed applications where hydrodynamic instability and misalignment are common problems. This bearing consists of a number of shoes mounted on pivots, with each shoe being a partial-arc bearing. The shoes adjust and follow the motions of the journal, ensuring inherent stability if the inertia of the shoes does not interfere with the adjustment ability of the bearing. The load direction may either pass between the two bottom shoes or it may pass through the pivot of the bottom shoe. The lubricant is incompressible (i.e., liquid), and the lubricant film is laminar. Figure 6.6 illustrates a tilting-pad bearing.

Bearings 87

Load Di

Journal

am

ete

rD

Bearing Clearance C

Arc L e n gth

Figure 6.5 Partial-arc bearing

Load

Di

Journal

am

et

Ψ

er

D

Bearing Housing

Tilting Shoe Clearance C

Figure 6.6 Tilting-pad bearing

Rolling Element or Antifriction Rolling element antifriction bearings are one of the most common types used in machinery. Antifriction bearings are based on rolling motion as opposed to the sliding motion of plain bearings. The use of rolling elements between rotating and stationary surfaces reduces the friction to a fraction of that resulting with the use of plain bearings. Use of rolling element bearings is determined by many factors, including load, speed, misalignment sensitivity, space limitations, and desire for precise shaft positioning. They support both radial and axial loads and are generally used in moderate- to high-speed applications. Unlike fluid-film plain bearings, rolling element bearings have the added ability to carry the full load of the rotor assembly at any speed. Where

88 Bearings

fluid-film bearings must have turning gear to support the rotor’s weight at low speeds, rolling element bearings can maintain the proper shaft centerline through the entire speed range of the machine.

Grade Classifications Rolling element bearings are available in either commercial- or precisiongrade classifications. Most commercial-grade bearings are made to nonspecific standards and are not manufactured to the same precise standards as precision-grade bearings. This limits the speeds at which they can operate efficiently, and given the brand of bearings may or may not be interchangeable. Precision bearings are used extensively on many machines such as pumps, air compressors, gear drives, electric motors, and gas turbines. The shape of the rolling elements determines the use of the bearing in machinery. Because of standardization in bearing envelope dimensions, precision bearings were once considered to be interchangeable, even if manufactured by different companies. It has been discovered, however, that interchanging bearings is a major cause of machinery failure and should be done with extreme caution.

Rolling Element Types There are two major classifications of rolling elements: ball and roller. Ball bearings function on point contact and are suited for higher speeds and lighter loads than roller bearings. Roller element bearings function on line contact and generally are more expensive than ball bearings, except for the larger sizes. Roller bearings carry heavy loads and handle shock more satisfactorily than ball bearings, but are more limited in speed. Figure 6.7 provides general guidelines to determine if a ball or roller bearing should be selected. This figure is based on a rated life of 30,000 hours. Although there are many types of rolling elements, each bearing design is based on a series of hardened rolling elements sandwiched between hardened inner and outer rings. The rings provide continuous tracks or races for the rollers or balls to roll in. Each ball or roller is separated from its neighbor by a separator cage or retainer, which properly spaces the rolling elements around the track and guides them through the load zone. Bearing size is usually given in terms of boundary dimensions: outside diameter, bore, and width.

Bearings 89

ed it

lim

Radial load, lb

Roller bearing

10,000 8 6 4 3 2

Ball or roller bearing

1000 8 6 4 3 2 100 8 6 4 3 2 10

Spe

6 4 3 2

Ball

2 3 4 6 100 2 3 4 6 1000 3 4 6 10,000 Shaft speed, rpm

Figure 6.7 Guide to selecting ball or roller bearings

Ball Bearings Common functional groupings of ball bearings are radial, thrust, and angular-contact bearings. Radial bearings carry a load in a direction perpendicular to the axis of rotation. Thrust bearings carry only thrust loads, a force parallel to the axis of rotation tending to cause endwise motion of the shaft. Angular-contact bearings support combined radial and thrust loads. These loads are illustrated in Figure 6.8. Another common classification of ball bearings is single-row (also referred to as Conrad or deep-groove bearing) and double-row.

Single-Row Types of single-row ball bearings are radial nonfilling slot bearings, radial filling slot bearings, angular contact bearings, and ball thrust bearings. Radial, Nonfilling Slot Bearings This ball bearing is often referred to as the Conrad-type or deep-groove bearing and is the most widely used of all ball bearings (and probably of

90 Bearings

(a) Radial load

(b) Thurst load

(c) Combination load

Figure 6.8 Three principal types of ball bearing loads

Figure 6.9 Single-row radial, nonfilling slot bearing

all antifriction bearings). It is available in many variations, with single or double shields or seals. They sustain combined radial and thrust loads, or thrust loads alone, in either direction—even at extremely high speeds. This bearing is not designed to be self-aligning; therefore, it is imperative that the shaft and the housing bore be accurately aligned (Figure 6.9). Figure 6.10 labels the parts of the Conrad antifriction ball bearing. This design is widely used and is versatile because the deep-grooved raceways permit the rotating balls to rapidly adjust to radial and thrust loadings, or a combination of these loadings. Radial, Filling Slot Bearing The geometry of this ball bearing is similar to the Conrad bearing, except for the filling slot. This slot allows more balls in the complement and thus can

Bearings 91

Width Corner radius Outer ring Shoulders Inner ring Corner radius Outside diameter Bore

Inner ring ball race

Separator

Face

Outer ring ball race

Figure 6.10 Conrad antifriction ball bearing parts

carry heavier radial loads. The bearing is assembled with as many balls that fit in the gap created by eccentrically displacing the inner ring. The balls are evenly spaced by a slight spreading of the rings and heat expansion of the outer ring. However, because of the filling slot, the thrust capacity in both directions is reduced. In combination with radial loads, this bearing design accomodates thrust of less than 60% of the radial load. Angular Contact Radial Thrust This ball bearing is designed to support radial loads combined with thrust loads, or heavy thrust loads (depending on the contact-angle magnitude). The outer ring is designed with one shoulder higher than the other, which allows it to accommodate thrust loads. The shoulder on the other side of the ring is just high enough to prevent the bearing from separating. This type of bearing is used for pure thrust load in one direction and is applied either in opposed pairs (duplex), or one at each end of the shaft. They can be mounted either face to face or back to back and in tandem for constant thrust in one direction. This bearing is designed for combination loads where the thrust component is greater than the capacity of single-row, deep-groove bearings. Axial deflection must be confined to very close tolerances.

92 Bearings

Ball-Thrust Bearing The ball-thrust bearing supports very high thrust loads in one direction only, but supports no radial loading. To operate successfully, this type of bearing must be at least moderately thrust-loaded at all times. It should not be operated at high speeds, since centrifugal force causes excessive loading of the outer edges of the races.

Double-Row Double-row ball bearings accommodate heavy radial and light thrust loads without increasing the outer diameter of the bearing. However, the doublerow bearing is approximately 60 to 80% wider than a comparable single-row bearing. The double-row bearing incorporates a filling slot, which requires the thrust load to be light. Figure 6.11 shows a double-row type ball bearing. This unit is, in effect, two single-row angular contact bearings built as a unit with the internal fit between balls and raceway fixed during assembly. As a result, fit and internal stiffness are not dependent upon mounting methods. These bearings usually have a known amount of internal preload, or compression, built in for maximum resistance to deflection under combined

Figure 6.11 Double row-type ball bearing

Bearings 93

Figure 6.12 Double-row internal self-aligning bearing

loads with thrust from either direction. As a result of this compression prior to external loading, the bearings are very effective for radial loads where bearing deflection must be minimized. Another double-row ball bearing is the internal self-aligning type, which is shown in Figure 6.12. It compensates for angular misalignment, which can be caused by errors in mounting, shaft deflection, misalignment, etc. This bearing supports moderate radial loads and limited thrust loads.

Roller As with plain and ball bearings, roller bearings also may be classified by their ability to support radial, thrust, and combination loads. Note that combination load-supporting roller bearings are not called angular-contact bearings as they are with ball bearings. For example, the taper-roller bearing is a combination load-carrying bearing by virtue of the shape of its rollers. Figure 6.13 shows the different types of roller elements used in these bearings. Roller elements are classified as cylindrical, barrel, spherical, and tapered. Note that barrel rollers are called needle rollers when they are less than 14 " in diameter and have a relatively high ratio of length to diameter.

94 Bearings

Spherical

Cylindrical

Needle

Tapered

Figure 6.13 Types of roller elements

Spherical

Cylindrical

Needle

Tapered

Figure 6.14 Cylindrical roller bearing

Cylindrical Cylindrical bearings have solid or helically wound hollow cylindrically shaped rollers, which have an approximate length-diameter ratio ranging from 1:1 to 1:3. They normally are used for heavy radial loads beyond the capacities of comparably sized radial ball bearings. Cylindrical bearings are especially useful for free axial movement of the shaft. The free ring may have a restraining flange to provide some restraint to endwise movement in one direction. Another configuration comes without a flange, which allows the bearing rings to be displaced axially. Either the rolls or the roller path on the races may be slightly crowned to prevent edge loading under slight shaft misalignment. Low friction makes this bearing type suitable for fairly high speeds. Figure 6.14 shows a typical cylindrical roller bearing. Figure 6.15 shows separable inner-ring cylindrical roller bearings. Figure 6.16 shows separable inner-ring cylindrical roller bearings with a different inner ring.

Bearings 95

Figure 6.15 Separable inner ring-type cylindrical roller bearings

Figure 6.16 Separable inner ring-type roller bearings with different inner ring

Figure 6.17 Separable inner ring-type cylindrical roller bearings with elimination of a retainer ring on one side The roller assembly in Figure 6.15 is located in the outer ring with retaining rings. The inner ring can be omitted and the roller operated on hardened ground shaft surfaces. The style in Figure 6.16 is similar to the one in Figure 6.15, except the rib on the inner ring is different. This prohibits the outer ring from moving in a direction toward the rib. Figure 6.17 shows separable inner ring-type cylindrical roller bearings with elimination of a retainer ring on one side.

96 Bearings

Figure 6.18 Needle bearings

The style shown in Figure 6.17 is similar to the two previous styles except for the elimination of a retainer ring on one side. It can carry small thrust loads in only one direction.

Needle-Type Cylindrical or Barrel Needle-type cylindrical bearings (Figure 6.18) incorporate rollers that are symmetrical, with a length at least four times their diameter. They are sometimes referred to as barrel rollers. These bearings are most useful where space is limited and thrust-load support is not required. They are available with or without an inner race. If a shaft takes the place of an inner race, it must be hardened and ground. The full-complement type is used for high loads and oscillating or slow speeds. The cage type should be used for rotational motion. They come in both single-row and double-row mountings. As with all cylindrical roller bearings, the single-row mounting type has a low thrust capacity, but angular mounting of rolls in the double-row type permits its use for combined axial and thrust loads.

Spherical Spherical bearings are usually furnished in a double-row mounting that is inherently self-aligning. Both rows of rollers have a common spherical outer raceway. The rollers are barrel-shaped with one end smaller to provide a small thrust to keep the rollers in contact with the center guide flange. This type of roller bearing has a high radial and moderate-to-heavy thrust load-carrying capacity. It maintains this capability with some degree of shaft and bearing housing misalignment. While their internal self-aligning feature is useful, care should be taken in specifying this type of bearing to compensate for misalignment. Figure 6.19 shows a typical spherical roller bearing assembly. Figure 6.20 shows a series of spherical roller bearings for a given shaft size.

Bearings 97

Figure 6.19 Spherical roller bearing assembly

239 230

240

231

241

222

232

213

223

4533

Figure 6.20 Series of spherical roller bearings for a given shaft size (available in several series)

Tapered Tapered bearings are used for heavy radial and thrust loads. They have straight tapered rollers, which are held in accurate alignment by means of a guide flange on the inner ring. Figure 6.21 shows a typical taperedroller bearing. Figure 6.22 shows necessary information to identify a taper-roller bearing. Figure 6.23 shows various types of tapered roller bearings. True rolling occurs because they are designed so all elements in the rolling surface and the raceways intersect at a common point on the axis. The basic characteristic of these bearings is that if the apexes of the tapered working surfaces of both rollers and races were extended, they would coincide on

98 Bearings

Figure 6.21 Tapered-roller bearing Bearing width Cup width

Cup front face radius

Cup back face radius

Cup outside diameter

Cone front face rib Cone front face Cone front face radius

Cup Roller Cone

Cup front face Cone back face rib Cone back face Cone back face radius

Cone width Cone bore

Cup back face Cage

Standout

Load pressure (effective) center distance

Figure 6.22 Information needed to identify a taper-roller bearing

Bearings 99

Single row

TS

FR

TSF

Double row

TDO

TNA

TDI

TNASWE

TDIT

TNASW

Four row

TQO

Figure 6.23 Various types of tapered roller bearings the bearing axis. Where maximum system rigidity is required, they can be adjusted for a preload. These bearings are separable.

Bearing Materials Because two contacting metal surfaces are in motion in bearing applications, material selection plays a crucial role in their life. Properties of the materials used in bearing construction determine the amount of sliding friction that occurs, a key factor affecting bearing life. When two similar metals are in contact without the presence of adequate lubrication, friction is generally

100 Bearings

high, and the surfaces will seize (i.e., weld) at relatively low pressures or surface loads. However, certain combinations of materials support substantial loads without seizing or welding as a result of their low frictional qualities. In most machinery, shafts are made of steel. Bearings are generally made of softer materials that have low frictional as well as sacrificial qualities when in contact with steel. A softer, sacrificial material is used for bearings because it is easier and cheaper to replace a worn bearing as opposed to a worn shaft. Common bearing materials are cast iron, bronze, and babbitt. Other less commonly used materials include wood, plastics, and other synthetics. There are several important characteristics to consider when specifying bearing materials, including: (1) strength or the ability to withstand loads without plastic deformation; (2) ability to permit embedding of grit or dirt particles that are present in the lubricant; (3) ability to elastically deform in order to permit load distribution over the full bearing surface; (4) ability to dissipate heat and prevent hot spots that might seize; and (5) corrosion resistance.

Plain As indicated above, dissimilar metals with low frictional characteristics are most suitable for plain bearing applications. With steel shafts, plain bearings made of bronze or babbitt are commonly used. Bronze is one of the harder bearing materials and is generally used for low speeds and heavy loads. A plain bearing may sometimes be made of a combination of materials. The outer portion may be constructed of bronze, steel, or iron to provide the strength needed to provide a load-carrying capability. The bearing may be lined with a softer material such as babbitt to provide the sacrificial capability needed to protect the shaft.

Rolling Element A specially developed steel alloy is used for an estimated 98% of all rolling element bearing uses. In certain special applications, however, materials such as glass, plastic, and other substances are sometimes used in rolling element construction. Bearing steel is a high-carbon chrome alloy with high hardenability and good toughness characteristics in the hardened and drawn state. All load-carrying members of most rolling contact bearings are made with this steel.

Bearings 101

Controlled procedures and practices are necessary to ensure specification of the proper alloy, maintain material cleanliness, and ensure freedom from defects—all of which affect bearing reliability. Alloying practices that conform to rigid specifications are required to reduce anomalies and inclusions that adversely affect a bearing’s useful life. Magnaflux inspections ensure that rolling elements are free from material defects and cracks. Light etching is used between rough and finish grinding processes to stop burning during heavy machining operations.

Lubrication It is critical to consider lubrication requirements when specifying bearings. Factors affecting lubricants include relatively high speeds, difficulty in performing relubrication, nonhorizontal shafts, and applications where leakage cannot be tolerated. This section briefly discusses lubrication mechanisms and techniques for bearings.

Plain Bearings In plain bearings, the lubricating fluid must be replenished to compensate for end leakage in order to maintain the bearings’ load-carrying capacity. Pressure lubrication from a pump- or gravity-fed tank, or automatic lubricating devices such as oil rings or oil disks, are provided in self-contained bearings. Another means of lubrication is to submerge the bearing (in particular, thrust bearings for vertical shafts) in an oil bath.

Lubricating Fluids Almost any process fluid may be used to lubricate plain bearings if parameters such as viscosity, corrosive action, toxicity, change in state (where a liquid is close to its boiling point), and in the case of a gaseous fluid, its compressibility, are appropriate for the application. Fluid-film journal and thrust bearings have run successfully, for example, on water, kerosene, gasoline, acid, liquid refrigerants, mercury, molten metals, and a wide variety of gases. Gases, however, lack the cooling and boundary-lubrication capabilities of most liquid lubricants. Therefore, the operation of self-acting gas bearings is restricted by start/stop friction and wear. If start/stop is performed under load, then the design is limited to about seven pounds per square

102 Bearings

inch (lb/in2 ) or 48 kilo-Newtons per square meter (kN/m2 ) on the projected bearing area, depending upon the choice of materials. In general, the materials used for these bearings are those of dry rubbing bearings (e.g., either a hard/hard combination such as ceramics with or without a molecular layer of boundary lubricant, or a hard/soft combination with a plastic surface). Externally pressurized gas journal bearings have the same principle of operation as hydrostatic liquid-lubricated bearings. Any clear gas can be used, but many of the design charts are based on air. There are three forms of external flow restrictors in use with these bearings: pocketed (simple) orifice, unpocketed (annular) orifice, and slot.

State of Lubrication Fluid or complete lubrication, the condition where the surfaces are completely separated by a fluid film, provides the lowest friction losses and prevents wear. The semifluid lubrication state exists between the journal and bearing when a load-carrying fluid film does not form to separate the surfaces. This occurs at comparatively low speed with intermittent or oscillating motion, heavy load, and insufficient oil supply to the bearing. Semifluid lubrication also may exist in thrust bearings with fixed parallel-thrust collars; guide bearings of machine tools; in bearings with plenty of lubrication that have a bent or misaligned shaft; or where the bearing surface has improperly arranged oil grooves. The coefficient of friction in such bearings may range from 0.02 to 0.08. In situations where the bearing is well lubricated, but the speed of rotation is very slow or the bearing is barely greasy, boundary lubrication takes place. In this situation, which occurs in bearings when the shaft is starting from rest, the coefficient of friction may vary from 0.08 to 0.14. A bearing may run completely dry in exceptional cases of design or with a complete failure of lubrication. Depending on the contacting surface materials, the coefficient of friction will be between 0.25 and 0.40.

Rolling Element Bearings Rolling element bearings also need a lubricant to meet or exceed their rated life. In the absence of high temperatures, however, excellent performance can be obtained with a very small quantity of lubricant. Excess lubricant causes excessive heating, which accelerates lubricant deterioration.

Bearings 103

Table 6.9 Ball-bearing grease relubrication intervals (hours of operation) Bearing speed, rpm Bearing bore, mm

5,000

3,600

1,750

1,000

200

10 20 30 40 50 60 70 80 90 100

8,700 5,500 4,000 2,800

12,000 8,000 6,000 4,500 3,500 2,600

25,000 17,000 13,000 11,000 9,300 8,000 6,700 5,700 4,800 4,000

44,000 30,000 24,000 20,000 18,000 16,000 14,000 12,000 11,000 10,000

220,000 150,000 127,000 111,000 97,000 88,000 81,000 75,000 70,000 66,000

Source: Theodore Baumeister, ed. Marks’ Standard Handbook for Mechanical Engineers, Eighth Edition. New York: McGraw-Hill, 1978.

The most popular type of lubrication is the sealed grease ball-bearing cartridge. Grease is commonly used for lubrication because of its convenience and minimum maintenance requirements. A high-quality lithium-based NLGI 2 grease is commonly used for temperatures up to 180◦ F (82◦ C). Grease must be replenished and relubrication intervals in hours of operation are dependent on temperature, speed, and bearing size. Table 6.9 is a general guide to the time after which it is advisable to add a small amount of grease. Some applications, however, cannot use the cartridge design, for example, when the operating environment is too hot for the seals. Another example is when minute leaks or the accumulation of traces of dirt at the lip seals cannot be tolerated (e.g., food processing machines). In these cases, bearings with specialized sealing and lubrication systems must be used. In applications involving high speed, oil lubrication is typically required. Table 6.10 is a general guide in selecting oil of the proper viscosity for these bearings. For applications involving high-speed shafts, bearing selection must take into account the inherent speed limitations of certain bearing designs, cooling needs, and lubrication issues such as churning and aeration suppression. A typical case is the effect of cage design and roller-end thrust-flange contact on the lubrication requirements in taper roller bearings. These design elements limit the speed and the thrust load that these

104 Bearings

Table 6.10 Oil lubrication viscosity (ISO identification numbers) Bearing speed, rpm Bearing bore, mm

10,000

3,600

1,800

600

50

4–7 10–20 25–45 50–70 75–90 100

68 32 10 7 3 3

150 68 32 22 10 7

220 150 68 68 22 22

220 150 150 68 68

460 320 320 220 220

Source: Theodore Baumeister, ed. Marks’ Standard Handbook for Mechanical Engineers, Eighth Edition. New York: McGraw-Hill, 1978.

bearings can endure. As a result, it is important always to refer to the bearing manufacturer’s instructions on load-carrying design and lubrication specifications.

Installation and General Handling Precautions Proper handling and installation practices are crucial to optimal bearing performance and life. In addition to standard handling and installation practices, the issue of emergency bearing substitutions is an area of critical importance. If substitute bearings are used as an emergency means of getting a machine back into production quickly, the substitution should be entered into the historical records for that machine. This documents the temporary change and avoids the possibility of the substitute bearing becoming a permanent replacement. This error can be extremely costly, particularly if the incorrectly specified bearing continually fails prematurely. It is important that an inferior substitute be removed as soon as possible and replaced with the originally specified bearing.

Plain Bearing Installation It is important to keep plain bearings from shifting sideways during installation and to ensure an axial position that does not interfere with shaft fillets. Both of these can be accomplished with a locating lug at the parting line.

Bearings 105

Less frequently used is a dowel in the housing, which protrudes partially into a mating hole in the bearing. The distance across the outside parting edges of a plain bearing are manufactured slightly greater than the housing bore diameter. During installation, a light force is necessary to snap it into place and, once installed, the bearing stays in place because of the pressure against the housing bore. It is necessary to prevent a bearing from spinning during operation, which can cause a catastrophic failure. Spinning is prevented by what is referred to as “crush.” Bearings are slightly longer circumferentially than their mating housings and upon installation this excess length is elastically deformed or “crushed.” This sets up a high radial contact pressure between the bearing and housing, which ensures good back contact for heat conduction and, in combination with the bore-to-bearing friction, prevents spinning. It is important that under no circumstances should the bearing parting lines be filed or otherwise altered to remove the crush.

Roller Bearing Installation A basic rule of rolling element bearing installation is that one ring must be mounted on its mating shaft or in its housing with an interference fit to prevent rotation. This is necessary because it is virtually impossible to prevent rotation by clamping the ring axially.

Mounting Hardware Bearings come as separate parts that require mounting hardware or as premounted units that are supplied with their own housings, adapters, and seals. Bearing Mountings Typical bearing mountings, which are shown in Figure 6.24, locate and hold the shaft axially and allow for thermal expansion and/or contraction of the shaft. Locating and holding the shaft axially is generally accomplished by clamping one of the bearings on the shaft so that all machine parts remain in proper relationship dimensionally. The inner ring is locked axially relative to the shaft by locating it between a shaft shoulder and some type of removable locking device once the inner ring has a tight fit. Typical removable locking devices are specially designed nuts, which are used for a through shaft, and clamp plates, which are commonly used when the bearing is mounted on the end of the shaft. For the locating or held bearing, the outer

106 Bearings

Locating

Nonlocating

Locating

(a)

Locating

Nonlocating (b)

Nonlocating

“Cross-locating”

(c)

(d)

Figure 6.24 Typical bearing mounting ring is clamped axially, usually between housing shoulders or end-cap pilots. With general types of cylindrical roller bearings, shaft expansion is absorbed internally simply by allowing one ring to move relative to the other [Figure 6.24(a) and 6.24(c), nonlocating positions]. The advantage of this type of mounting is that both inner and outer rings may have a tight fit, which is desirable or even mandatory if significant vibration and/or imbalance exists in addition to the applied load. Premounted Bearing Premounted bearings, referred to as pillow-block and flanged-housing mountings, are of considerable importance to millwrights. They are particularly adaptable to “line-shafting” applications, which are a series of ball and roller bearings supplied with their own housings, adapters, and seals. Premounted bearings come with a wide variety of flange mountings, which permit them to be located on faces parallel or perpendicular to the shaft

Bearings 107

Figure 6.25 Typical pillow block

Figure 6.26 Flanged bearing unit

axis. Figure 6.25 shows a typical pillow block. Figure 6.26 shows a flanged bearing unit. Inner races can be mounted directly on ground shafts or can be adaptermounted to “drill-rod” or to commercial-shafting. For installations sensitive to imbalance and vibration, the use of accurately ground shaft seats is recommended. Most pillow-block designs incorporate self-aligning bearing types and do not require the precision mountings utilized with other bearing installations.

Mounting Techniques When mounting or dismounting a roller bearing, the most important thing to remember is to apply the mounting or dismounting force to the side face of the ring with the interference fit. This force should not pass from one

108 Bearings

ring to the other through the ball or roller set, as internal damage can easily occur. Mounting tapered-bore bearings can be accomplished simply by tightening the locknut or clamping plate. This locates it on the shaft until the bearing is forced the proper distance up the taper. This technique requires a significant amount of force, particularly for large bearings. Cold Mounting Cold mounting, or force fitting a bearing onto a shaft or into a housing, is appropriate for all small bearings (i.e., 4" bore and smaller). The force, however, must be applied as uniformly as possible around the side face of the bearing and to the ring to be press-fitted. Mounting fixtures, such as a simple piece of tubing of appropriate size and a flat plate, should be used. It is not appropriate to use a drift and hammer to force the bearing on, which will cause the bearing to cock. It is possible to apply force by striking the plate with a hammer or by an arbor press. However, before forcing the bearing on the shaft, a coat of light oil should be applied to the bearing seat on the shaft and the bearing bores. All sealed and shielded ball bearings should be cold mounted in this manner. Temperature Mounting The simplest way to mount any open straight-bore bearing regardless of its size is temperature mounting, which entails heating the entire bearing, pushing it on its seat, and holding it in place until it cools enough to grip the shaft. The housing may be heated if practical for tight outside-diameter fits; however, temperatures should not exceed 250◦ F. If heating of the housing is not practical, the bearing may be cooled with dry ice. The risk of cooling is that if the ambient conditions are humid, moisture is introduced and there is a potential for corrosion in the future. Acceptable ways of heating bearings are by hot plate, temperature-controlled oven, induction heaters, and hot-oil bath. With the hot plate method, the bearing is simply laid on the plate until it reaches the approved temperature, using a pyrometer or Tempilstik to make certain it is not overheated. Difficulty in controlling the temperature is the major disadvantage of this method. When using a temperature-controlled oven, the bearings should be left in the oven long enough to heat thoroughly, but they should never be left overnight.

Bearings 109

The use of induction heaters is a quick method of heating bearings. However, some method of measuring the ring temperature (e.g., pyrometer or a Tempilstik) must be used or damage to the bearing may occur. Note that bearings must be demagnetized after the use of this method. The use of a hot-oil bath is the most practical means of heating larger bearings. Disadvantages are that the temperature of the oil is hard to control and may ignite or overheat the bearing. The use of a soluble oil and water mixture (10 to 15% oil) can eliminate these problems and still attain a boiling temperature of 210◦ F. The bearing should be kept off the bottom of the container by a grate or screen located several inches off the bottom. This is important to allow contaminants to sink to the bottom of the container and away from the bearing.

Dismounting Commercially available bearing pullers allow rolling element bearings to be dismounted from their seats without damage. When removing a bearing, force should be applied to the ring with the tight fit, although sometimes it is necessary to use supplementary plates or fixtures. An arbor press is equally effective at removing smaller bearings as well as mounting them.

Ball Installation Figure 6.27 shows the ball installation procedure for roller bearings. The designed load carrying capacity of Conrad-type bearings is determined by the number of balls that can be installed between the rings. Ball installation is accomplished by the following procedure: ●

Slip the inner ring slightly to one side;

●

Insert balls into the gap, which centers the inner ring as the balls are positioned between the rings;

●

Place stamped retainer rings on either side of the balls before riveting together. This positions the balls equidistant around the bearing.

General Roller-Element Bearing Handling Precautions In order for rolling element bearings to achieve their design life and perform with no abnormal noise, temperature rise, or shaft excursions, the following precautions should be taken: ●

Always select the best bearing design for the application and not the cheapest. The cost of the original bearing is usually small by comparison

110 Bearings

1. The inner ring is moved to one side

2. Balls are installed in the gap

3. The inner ring is centered to the balls are equally positioned in place

4. A retainer is installed

Figure 6.27 Ball installation procedures to the costs of replacement components and the downtime in production when premature bearing failure occurs because an inappropriate bearing was used. ●

If in doubt about bearings and their uses, consult the manufacturer’s representative and the product literature.

●

Bearings should always be handled with great care. Never ignore the handling and installation instructions from the manufacturer.

●

Always work with clean hands, clean tools, and the cleanest environment available.

●

Never wash or wipe bearings prior to installation unless the instructions specifically state that this should be done. Exceptions to this rule are when oil-mist lubrication is to be used and the slushing compound has hardened in storage or is blocking lubrication holes in the bearing rings. In this situation, it is best to clean the bearing with kerosene or other appropriate petroleum-based solvent. The other exception is if the slushing compound has been contaminated with dirt or foreign matter before mounting.

●

Keep new bearings in their greased paper wrappings until they are ready to install. Place unwrapped bearings on clean paper or lint-free cloth if they cannot be kept in their original containers. Wrap bearings in clean, oil-proof paper when not in use.

●

Never use wooden mallets, brittle or chipped tools, or dirty fixtures and tools when bearings are being installed.

Bearings 111

●

Do not spin bearings (particularly dirty ones) with compressed service air.

●

Avoid scratching or nicking bearing surfaces. Care must be taken when polishing bearings with emery cloth to avoid scratching.

●

Never strike or press on race flanges.

●

Always use adapters for mounting that ensure uniform steady pressure rather than hammering on a drift or sleeve. Never use brass or bronze drifts to install bearings as these materials chip very easily into minute particles that will quickly damage a bearing.

●

Avoid cocking bearings onto shafts during installation.

●

Always inspect the mounting surface on the shaft and housing to insure that there are no burrs or defects.

●

When bearings are being removed, clean housings and shafts before exposing the bearings.

●

Dirt is abrasive and detrimental to the designed life span of bearings.

●

Always treat used bearings as if they are new, especially if they are to be reused.

●

Protect dismantled bearings from moisture and dirt.

●

Use clean filtered, water-free Stoddard’s solvent or flushing oil to clean bearings.

●

When heating is used to mount bearings onto shafts, follow the manufacturer’s instructions.

●

When assembling and mounting bearings onto shafts, never strike the outer race or press on it to force the inner race. Apply the pressure on the inner race only. When dismantling, follow the same procedure.

●

Never press, strike, or otherwise force the seal or shield on factory-sealed bearings.

Bearing Failures, Deficiencies, and Their Causes The general classifications of failures and deficiencies requiring bearing removal are overheating, vibration, turning on the shaft, binding of the

112 Bearings

shaft, noise during operation, and lubricant leakage. Table 6.11 is a troubleshooting guide that lists the common causes for each of these failures and deficiencies. As indicated by the causes of failure listed, bearing failures are rarely caused by the bearing itself. Many abnormal vibrations generated by actual bearing problems are the result of improper sizing of the bearing liner or improper lubrication. However, numerous machine and process-related problems generate abnormal vibration spectra in bearing data. The primary contributors to abnormal bearing signatures are: (1) imbalance, (2) misalignment, (3) rotor instability, (4) excessive or abnormal loads, and (5) mechanical looseness. Defective bearings that leave the manufacturer are very rare, and it is estimated that defective bearings contribute to only 2% of total failures. The failure is invariably linked to symptoms of misalignment, imbalance, resonance, and lubrication—or the lack of it. Most of the problems that occur result from the following reasons: dirt, shipping damage, storage and handling, poor fit resulting in installation damage, wrong type of bearing design, overloading, improper lubrication practices, misalignment, bent shaft, imbalance, resonance, and soft foot. Anyone of these conditions will eventually destroy a bearing—two or more of these problems can result in disaster! Although most industrial machine designers provide adequate bearings for their equipment, there are some cases in which bearings are improperly designed, manufactured, or installed at the factory. Usually, however, the trouble is caused by one or more of the following reasons: (1) improper on-site bearing selection and/or installation, (2) incorrect grooving, (3) unsuitable surface finish, (4) insufficient clearance, (5) faulty relining practices, (6) operating conditions, (7) excessive operating temperature, (8) contaminated oil supply, and (9) oil-film instability.

Improper Bearing Selection and/or Installation There are several things to consider when selecting and installing bearings, including the issue of interchangeability, materials of construction, and damage that might have occurred during shipping, storage, and handling.

Interchangeability Because of the standardization in envelope dimensions, precision bearings were once regarded as interchangeable among manufacturers.

Table 6.11 Troubleshooting guide Binding of the shaft

Noisy bearing

Lubricant leakage

Growth of race due to overheating Fretting wear

Lubricant breakdown

Lubrication breakdown

Overfilling of lubricant

Contamination by abrasive or corrosive materials Housing distortion or out-of-round pinching bearing Uneven shimming of housing with loss of clearance Tight rubbing seals

Inadequate lubrication

Grease churning due to too soft a consistency

Pinched bearing

Grease deterioration due to excessive operating temperature

Contamination

Operating beyond grease life

Seal rubbing

Seal wear

Preloaded bearings

Bearing slipping on shaft or in housing

Wrong shaft attitude (bearing seals designed for horizontal mounting only)

Vibration

Inadequate or insufficient lubrication Excessive lubrication

Dirt or chips in bearing

Grease liquifaction or aeration

Rotor unbalance

Improper initial fit

Oil foaming

Out-of-round shaft

Excessive shaft deflection

Abrasion or corrosion due to contaminants Housing distortion due to warping or out-of-round

Race misalignment

Initial coarse finish on shaft Seal rub on inner race

Fatigued race or rolling elements

Housing resonance

Continued

Bearings 113

Turning on the shaft

Overheating

Table 6.11 continued

Noisy bearing

Lubricant leakage

Cage wear

Cocked races

Flatted roller or ball

Seal failure

Flats on races or rolling elements

Loss of clearance due to excessive adapter tightening Thermal shaft expansion

Brinelling due to assembly abuse, handling, or shock loads

Clogged breather

Variation in size of rolling elements

Oil foaming due to churning or air flow through housing Gasket (O-ring) failure or misapplication Porous housing or closure Lubricator set at the wrong flow rate

Vibration

Seal rubbing or failure Inadequate or blocked scavenge oil passages

Inadequate bearing clearance or bearing preload Race turning

Race turning

Cage wear

Excessive clearance Corrosion False brinelling or indentation of races Electrical arcing Mixed rolling element diameters Out-of-square rolling paths in races

Source: Integrated Systems Inc.

Turning on the shaft

Out-of-round or lobular shaft Housing bore waviness Chips or scores under bearing seat

114 Bearings

Binding of the shaft

Overheating

Bearings 115

This interchangeability has since been considered a major cause of failures in machinery, and the practice should be used with extreme caution. Most of the problems with interchangeability stem from selecting and replacing bearings based only on bore size and outside diameters. Often, very little consideration is paid to the number of rolling elements contained in the bearings. This can seriously affect the operational frequency vibrations of the bearing and may generate destructive resonance in the host machine or adjacent machines. More bearings are destroyed during their installation than fail in operation. Installation with a heavy hammer is the usual method in many plants. Heating the bearing with an oxy-acetylene burner is another classical method. However, the bearing does not stand a chance of reaching its life expectancy when either of these installation practices are used. The bearing manufacturer’s installation instructions should always be followed.

Shipping Damage Bearings and the machinery containing them should be properly packaged to avoid damage during shipping. However, many installed bearings are exposed to vibrations, bending, and massive shock loadings through bad handling practices during shipping. It has been estimated that approximately 40% of newly received machines have “bad” bearings. Because of this, all new machinery should be thoroughly inspected for defects before installation. Acceptance criteria should include guidelines that clearly define acceptable design/operational specifications. This practice pays big dividends by increasing productivity and decreasing unscheduled downtime.

Storage and Handling Storeroom and other appropriate personnel must be made aware of the potential havoc they can cause by their mishandling of bearings. Bearing failure often starts in the storeroom rather than the machinery. Premature opening of packages containing bearings should be avoided whenever possible. If packages must be opened for inspection, they should be protected from exposure to harmful dirt sources and then resealed in the original wrappings. The bearing should never be dropped or bumped as this can cause shock loading on the bearing surface.

116 Bearings

Incorrect Placement of Oil Grooves Incorrectly placed oil grooves can cause bearing failure. Locating the grooves in high-pressure areas causes them to act as pressure-relief passages. This interferes with the formation of the hydrodynamic film, resulting in reduced load-carrying capability.

Unsuitable Surface Finish Smooth surface finishes on both the shaft and the bearing are important to prevent surface variations from penetrating the oil film. Rough surfaces can cause scoring, overheating, and bearing failure. The smoother the finishes, the closer the shaft may approach the bearing without danger of surface contact. Although important in all bearing applications, surface finish is critical with the use of harder bearing materials such as bronze.

Insufficient Clearance There must be sufficient clearance between the journal and bearing in order to allow an oil film to form. An average diametral clearance of 0.001 inches per inch of shaft diameter is often used. This value may be adjusted depending on the type of bearing material, the load, speed, and the accuracy of the shaft position desired.

Faulty Relining Faulty relining occurs primarily with babitted bearings rather than precision machine-made inserts. Babbitted bearings are fabricated by a pouring process that should be performed under carefully controlled conditions. Some reasons for faulty relining are: (1) improper preparation of the bonding surface, (2) poor pouring technique, (3) contamination of babbitt, and (4) pouring bearing to size with journal in place.

Operating Conditions Abnormal operating conditions or neglect of necessary maintenance precautions cause most bearing failures. Bearings may experience premature and/or catastrophic failure on machines that are operated heavily loaded, speeded up, or being used for a purpose not appropriate for the system design. Improper use of lubricants can also result in bearing failure. Some typical causes of premature failure include: (1) excessive operating temperatures, (2) foreign material in the lubricant supply, (3) corrosion, (4) material fatigue, and (5) use of unsuitable lubricants.

Bearings 117

Excessive Temperatures Excessive temperatures affect the strength, hardness, and life of bearing materials. Lower temperatures are required for thick babbitt liners than for thin precision babbitt inserts. Not only do high temperatures affect bearing materials, they also reduce the viscosity of the lubricant and affect the thickness of the film, which affects the bearing’s load-carrying capacity. In addition, high temperatures result in more rapid oxidation of the lubricating oil, which can result in unsatisfactory performance.

Dirt and Contamination in Oil Supply Dirt is one of the biggest culprits in the demise of bearings. Dirt makes its appearance in bearings in many subtle ways, and it can be introduced by bad work habits. It also can be introduced through lubricants that have been exposed to dirt, a problem that is responsible for approximately half of bearing failures throughout the industry. To combat this problem, soft materials such as babbit are used when it is known that a bearing will be exposed to abrasive materials. Babbitt metal embeds hard particles, which protects the shaft against abrasion. When harder materials are used in the presence of abrasives, scoring and galling occurs as a result of abrasives caught between the journal and bearing. In addition to the use of softer bearing materials for applications where abrasives may potentially be present, it is important to properly maintain filters and breathers, which should regularly be examined. In order to avoid oil supply contamination, foreign material that collects at the bottom of the bearing sump should be removed on a regular basis.

Oil-Film Instability The primary vibration frequency components associated with fluid-film bearings problems are in fact displays of turbulent or nonuniform oil film. Such instability problems are classified as either oil whirl or oil whip depending on the severity of the instability. Machine-trains that use sleeve bearings are designed based on the assumption that rotating elements and shafts operate in a balanced and, therefore, centered position. Under this assumption, the machine-train shaft will operate with an even, concentric oil film between the shaft and sleeve bearing. For a normal machine, this assumption is valid after the rotating element has achieved equilibrium. When the forces associated with rotation are

118 Bearings

Destabilizing component of bearing force

Lower pressure bearing fluid

Wh

n tio ta

Centrifugal force

irl Ro

Support component of bearing force

ng ari Be id e flu ssur t pre ultan res

Higher pressure bearing fluid

Undeflected shaft axis

Elastic restoring force

External damping

En flu tra flo id ined dir w ec tio n

Figure 6.28 Oil whirl, oil whip

in balance, the rotating element will center the shaft within the bearing. However, several problems directly affect this self-centering operation. First, the machine-train must be at designed operating speed and load to achieve equilibrium. Second, any imbalance or abnormal operation limits the machine-train’s ability to center itself within the bearing. A typical example is a steam turbine. A turbine must be supported by auxiliary running gear during startup or shutdown to prevent damage to the sleeve bearings. The lower speeds during the startup and shutdown phase

Bearings 119

of operation prevent the self-centering ability of the rotating element. Once the turbine has achieved full speed and load, the rotating element and shaft should operate without assistance in the center of the sleeve bearings. Oil Whirl In an abnormal mode of operation, the rotating shaft may not hold the centerline of the sleeve bearing. When this happens, an instability called oil whirl occurs. Oil whirl is an imbalance in the hydraulic forces within a sleeve bearing. Under normal operation, the hydraulic forces such as velocity and pressure are balanced. If the rotating shaft is offset from the true centerline of the bearing, instability occurs. As Figure 6.28 illustrates, a restriction is created by the offset. This restriction creates a high pressure and another force vector in the direction of rotation. Oil whirl accelerates the wear and failure of the bearing and bearing support structure. Oil Whip The most severe damage results if the oil whirl is allowed to degrade into oil whip. Oil whip occurs when the clearance between the rotating shaft and sleeve bearing is allowed to close to a point approaching actual metal-to-metal contact. When the clearance between the shaft and bearing approaches contact, the oil film is no longer free to flow between the shaft and bearing. As a result, the oil film is forced to change directions. When this occurs, the high-pressure area created in the region behind the shaft is greatly increased. This vortex of oil increases the abnormal force vector created by the offset and rotational force to the point that metal-to-metal contact between the shaft and bearing occurs. In almost all instances where oil whip is allowed, severe damage to the sleeve bearing occurs.

7

Chain Drives

“Only Permanent Repairs Made Here” Introduction Chain drives are an important part of a conveyor system. They are used to transmit needed power from the drive unit to a portion of the conveyor system. This chapter will cover: 1 Various types of chains that are used to transmit power in a conveyor system. 2

The advantages and disadvantages of using chain drives.

3

The correct installation procedure for chain drives.

4

How to maintain chain drives.

5

How to calculate speeds and ratios that will enable you to make corrections or adjustments to conveyor speeds.

6

How to determine chain length and sprocket sizes when making speed adjustments.

Chain Drives Chain drives are used to transmit power between a drive unit and a driven unit. For example, if we have a gearbox and a contact roll on a conveyor, we need a way to transmit the power from the gearbox to the roll. This can be done easily and efficiently with a chain drive unit. Chain drives can consist of one or multiple strand chains, depending on the load that the unit must transmit. The chains need to be the matched with the sprocket type, and they must be tight enough to prevent slippage. Chain is sized by the pitch or the center-to-center distance between the pins. This is done in 18 " increments, and the pitch number is found on the side bars. Examples of the different chain and sprocket sizes can be seen in Figures 7.1 and 7.2.

Chain Drives 121

35

3/8"

Figure 7.1 Chain size

40

4/8"

Figure 7.2 Chain size Sometimes chains are linked to form two multistrand chains. The number designation for this chain would have the same pitch number as standard chain, but the pitch would be followed by the number of strands (80-4).

Chain Selection Plain or Detachable-Link Chain Plain chains are usually used in slow speed applications like conveyors. They are rugged, designed to carry heavy loads, and when properly maintained

122 Chain Drives

can offer years of reliable service. They are made up of a series of detachable links that do not have rollers. The problem is that if the direction of the chain is reversed, the chain can come apart. When replacing a motor, the rotation of the coupling must be the same before you connect the coupling to the driven unit.

Roller Chain Roller chains are made up of roller links that are joined with pin links. The links are made up of two side bars, two rollers, and two bushings. The roller reduces the friction between the chain and the sprocket, thereby increasing the life of the unit. Roller chains can operate at faster speeds than plain chains, and properly maintained, they will offer years of reliable service. Some roller chains come with a double pitch, meaning that the pitch is double that of a standard chain, but the width and roller size remains the same. Double-pitch chain can be used on standard sprockets, but doublepitch sprockets are also available. The main advantage to the double-pitch chain is that it is cheaper than the standard pitch chain. So, they are often used for applications that require slow speeds, as in for lifting pieces of equipment in a hot press application.

Sprockets Sprockets are fabricated from a variety of materials; this would depend upon the application of the drive. Large fabricated steel sprockets are manufactured with holes to reduce the weight of the sprocket on the equipment. Because roller chain drives sometimes have restricted spaces for their installation or mounting, the hubs are made in several different styles. See Figure 7.3. Type A sprockets are flat and have no hub at all. They are usually mounted on flanges or hubs of the device that they are driving. This is accomplished through a series of holes that are either plain or tapered. Type B sprocket hubs are flush on one side and extend slightly on the other side. The hub is extended to one side to allow the sprocket to be fitted close to the machinery that it is being mounted on. This eliminates a large overhung load on the bearings of the equipment. Type C sprockets are extended on both sides of the plate surface. They are usually used on the driven sprocket where the pitch diameter is larger and

Chain Drives 123

Hub Classification

A

B

C

D

Figure 7.3 Types of sprocket hubs where there is more weight to support on the shaft. Remember this the larger the load is, the larger the hub should be. Type D sprockets use an A sprocket mounted on a solid or split hub. The type A sprocket is split and bolted to the hub. This is done for ease of removal and not practicality. It allows the speed ratio to be changed easily by simply unbolting the sprocket and changing it without having the remove bearings or other equipment.

Chain Installation When the proper procedures are followed for installing chains, they will yield years of trouble-free service. Use the following procedure to perform this task: 1

The shafts must be parallel, or the life of the chain will be shortened. The first step is to level the shafts. This is done by placing a level on each of the shafts, then shimming the low side until the shaft is level. See Figure 7.4.

2

The next step is to make sure that the shafts are parallel. This is done by measuring at different points on the shaft and adjusting the shafts until

124 Chain Drives

they are an equal distance apart. Make sure that the shafts are pulled in as close as possible before performing this procedure. The jacking bolts can be used to move the shafts apart evenly after the chain is installed. See Figure 7.5. 3

Before installing a set of used sprockets, verify the size and condition of the sprockets.

4

Install the sprockets on the shafts following the manufacturer’s recommendations. Locate and install the first sprocket, then use a straightedge or a string to line the other one up with the one previously installed.

Figure 7.4 Alignment

25"

Figure 7.5 Alignment

25"

Chain Drives 125

5

Install the chain on the sprockets, then begin increasing the distance between the sprockets by turning the jacking bolts; do this until the chain is snug but not tight. To set the proper chain sag, deflect the chain 1 " per foot of span between the shafts. Use a string or straightedge and 4 place it across the top of the chain. Then push down on the chain just enough to remove the slack. Use a tape measure to measure the amount of sag. See Figure 7.6.

6

Do a final check for parallel alignment. Remember: the closer the alignment, the longer the chain will run. See Figure 7.7.

5" 4" 3" 2" 1"

24"

Figure 7.6 Tensioning

Figure 7.7 Final alignment

126 Chain Drives

Power Train Formulas Shaft Speed The size of the sprockets in a chain drive system determines the speed relationship between the drive and driven sprockets. For example, if the drive sprocket has the same size sprocket as the driven, then the speed will be equal. See Figure 7.8. If we change the size of driven sprocket, then the speed of the shaft will also change. If we know what the speed of the electric motor is, and the size of the sprockets, we can calculate the speed of the driven shaft by using the following formula (see Figure 7.9): Driven shaft rpm =

Drive sprocket # teeth × drive shaft rpm Driven sprocket # teeth

Driven

Drive

6T

6T

1800 rpm

1800 rpm

Figure 7.8 Ratio Driven

Drive

12T

6T

____ rpm

1800 rpm

Figure 7.9 Speed ratio example

Chain Drives 127

6 × 1800 12 6 × 1800 900 = 12

Driven shaft rpm =

Now we understand how changing the size of a sprocket will also change the shaft speed. Knowing this, we could also assume that to change the shaft rpm we must change the sprocket size. The problem is how do we know the exact size sprocket that we need to reach the desired speed? Use the same formula that was used to calculate shaft speed, only switch the location of the driven shaft speed and the driven sprocket size: Driven shaft rpm =

Drive sprocket # teeth × drive shaft rpm Driven sprocket # teeth

Let’s change the problem to look like this: Driven sprocket teeth =

Drive sprocket # teeth × drive shaft rpm Driven shaft rpm

Let’s say that we have a problem similar to the ones that we just did, but we want to change the shaft speed of the driven unit. If we know the speed we are looking for, we can use the formula above to calculate the sprocket size required. Let’s change the speed of the driven shaft to 900 rpm (see Figure 7.10):

Driven

Drive

12T

6T

____ rpm

1800 rpm

Figure 7.10 Sprocket calculations

128 Chain Drives

Driven shaft rpm =

6 × 1800 900

12 =

6 × 1800 900

Chain Length Many times when a mechanic has to change out chains there is no way of knowing how long the chain should be. One way is to lay the new chain down beside the old chain, but remember that the old chain has been stretched. Or, maybe you are installing a new drive and you want to have the chain made up before you install it. So what do you do? One method is to take a tape measure and wrap it around the sprockets to get the chain length. However, this is not a very accurate way to determine the length. Instead, let’s take a couple of measurements, then use a simple formula to calculate the actual length that is needed. First, move the sprockets together until they are as close as the adjustments will allow. Then move the motor or drive out 14 of its travel. Now we are ready to take our measurements. The following information is needed for an equation to find the chain length: 1

Number of teeth on the drive sprocket.

2

Center-to-center distance between the shafts.

3

The chain pitch in inches.

Now use the following formula to solve the equation (see Figure 7.11): Chain length =

# teeth drive × pitch 2 # teeth driven × pitch + 2 + center to center × 2

Chain Drives 129

Driven

Drive

40 chain

12T

6T

35"

Figure 7.11 Chain size calculations Use the formula above to find the chain length. 6×. 5 12×. 5 + 35" × 2 + 2 2 6×. 5 12×. 5 + 35" × 2 + 74. 5" = 2 2

Chain length =

Multiple Sprockets When calculating multiple sprocket systems, think of each set of sprockets as a two-sprocket system.

Chain Speed In order to calculate the speed of a chain in feet per minute (FPM), we need the following information: 1

The number of teeth on the sprocket.

2

The shaft rpm of the sprocket.

3

The pitch of the chain in inches.

130 Chain Drives

Driven

12T

Drive

40 chain

6T

900 rpm

1800 rpm

Figure 7.12 Speed calculations

With this information we can use the following formula: FPM =

# teeth × pitch × rpm 12

Use this formula to find the speed of the following chain (see Figure 7.12):

FPM =

# teeth × pitch × rpm 12

450 =

6T × .5" × 1800 12

Chain drives are used to transmit power between a drive unit and a driven unit. For example, if we have a gearbox and a contact roll on a conveyor, we need a way to transmit the power from the gearbox to the roll. This can be done easily and efficiently with a chain drive unit. Chain drives can consist of one or multiple strand chains, depending on the load that the unit must transmit. The chains need to be matched with the sprocket type, and they must be tight enough to prevent slippage. An effective preventive maintenance program will provide extended life to a chain drive system, and through proper corrective maintenance procedures, we can prevent premature failures.

Chain Drives 131

Preventive Maintenance Procedures Inspection (risk of failures for not following the procedures below is noted along with a rating): LOW: minimal risk/low chance of failure; MEDIUM: failure is possible but equipment not operation to specification is highly probable; HIGH: failure will happen prematurely. ●

Inspect a chain for wear by inspecting the links for worn bushings. If worn bushings are noted, write a corrective maintenance work order so that the replacement can be planned and scheduled at a later time.

Risk if the procedure is not followed: HIGH. Chain breakage will occur. ●

Lubricate chain with lightweight oil recommended by chain manufacturer. (Ask your chain supplier to visit your site and make recommendations based on documentation they can present to you.)

Risk if the procedure is not followed: HIGH. Chain breakage will occur. ●

Check chain sag. Measure the chain sag using a straight edge or string and measure the specifications noted on this PM task. (The chain sag specification can be provided by your chain supplier, or you can use the procedure noted earlier in this chapter.) WARNING: The specification must be noted on the PM procedure.

●

Set tension, and make a note at the bottom of the PM work order, if a deficiency is noted.

Risk if the procedure is not followed: MEDIUM. Sprocket and chain wear will accelerate, thus causing equipment stoppage. ●

Inspect sprockets for worn teeth and abnormal wear on the sides of the sprockets. (The question is: Can the sprockets and chain last for two more weeks without equipment stoppage?) If the sprockets and chain can last two weeks then write a corrective maintenance work order in order for this job to be planned and scheduled with the correct parts. If the sprocket cannot last two weeks, then change all sprockets and the chain. Set and check sheave and chain alignment and tension. WARNING: When

132 Chain Drives

changing a sprocket, all sprockets, and the chain, should be changed because the difference between a worn and new sprocket in pitch diameter can be extreme, thus causing premature failure of the sprockets and chain. Risk if the procedure is not followed: High. Worn sprockets are an indication of the equipment being in a failure mode. Action must be taken.

8

Compressors

A compressor is a machine that is used to increase the pressure of a gas or vapor. They can be grouped into two major classifications: centrifugal and positive displacement. This section provides a general discussion of these types of compressors.

Centrifugal In general, the centrifugal designation is used when the gas flow is radial and the energy transfer is predominantly due to a change in the centrifugal forces acting on the gas. The force utilized by the centrifugal compressor is the same as that utilized by centrifugal pumps. In a centrifugal compressor, air or gas at atmospheric pressure enters the eye of the impeller. As the impeller rotates, the gas is accelerated by the rotating element within the confined space that is created by the volute of the compressor’s casing. The gas is compressed as more gas is forced into the volute by the impeller blades. The pressure of the gas increases as it is pushed through the reduced free space within the volute. As in centrifugal pumps, there may be several stages to a centrifugal air compressor. In these multistage units, a progressively higher pressure is produced by each stage of compression.

Configuration The actual dynamics of centrifugal compressors are determined by their design. Common designs are: overhung or cantilever, centerline, and bullgear.

Overhung or Cantilever The cantilever design is more susceptible to process instability than centerline centrifugal compressors. Figure 8.1 illustrates a typical cantilever design. The overhung design of the rotor (i.e., no outboard bearing) increases the potential for radical shaft deflection. Any variation in laminar flow, volume,

134 Compressors

Figure 8.1 Cantilever centrifugal compressor is susceptible to instability or load of the inlet or discharge gas forces the shaft to bend or deflect from its true centerline. As a result, the mode shape of the shaft must be monitored closely.

Centerline Centerline designs, such as horizontal and vertical split-case, are more stable over a wider operating range, but should not be operated in a variabledemand system. Figure 8.2 illustrates the normal airflow pattern through a horizontal split-case compressor. Inlet air enters the first stage of the compressor, where pressure and velocity increases occur. The partially compressed air is routed to the second stage where the velocity and pressure are increased further. Adding additional stages until the desired final discharge pressure is achieved can continue this process. Two factors are critical to the operation of these compressors: impeller configuration and laminar flow, which must be maintained through all of the stages. The impeller configuration has a major impact on stability and operating envelope. There are two impeller configurations: in-line and back-to-back, or opposed. With the in-line design, all impellers face in the same direction. With the opposed design, impeller direction is reversed in adjacent stages.

Compressors 135

Figure 8.2 Airflow through a centerline centrifugal compressor

To discharge

Balancing line to suction

Shaft seal

Balancing piston

Figure 8.3 Balancing piston resists axial thrust from the in-line impeller design of a centerline centrifugal compressor

In-Line A compressor with all impellers facing in the same direction generates substantial axial forces. The axial pressures generated by each impeller for all the stages are additive. As a result, massive axial loads are transmitted to the fixed bearing. Because of this load, most of these compressors use either a Kingsbury thrust bearing or a balancing piston to resist axial thrusting. Figure 8.3 illustrates a typical balancing piston.

136 Compressors

All compressors that use in-line impellers must be monitored closely for axial thrusting. If the compressor is subjected to frequent or constant unloading, the axial clearance will increase due to this thrusting cycle. Ultimately, this frequent thrust loading will lead to catastrophic failure of the compressor.

Opposed By reversing the direction of alternating impellers, the axial forces generated by each impeller or stage can be minimized. In effect, the opposed impellers tend to cancel the axial forces generated by the preceding stage. This design is more stable and should not generate measurable axial thrusting. This allows these units to contain a normal float and fixed rolling-element bearing.

Bullgear The bullgear design uses a direct-driven helical gear to transmit power from the primary driver to a series of pinion-gear-driven impellers that are located around the circumference of the bullgear. Figure 8.4 illustrates a typical bullgear compressor layout. First-stage diffuser

First-stage intercooler

Condensate separator

First-stage rotor Second-stage inlet First-stage inlet

Bull gear Fourth-stage rotor

Third-stage inlet Fourth-stage inlet

Aftercooler Discharge

Figure 8.4 Bullgear centrifugal compressor

Compressors 137

The pinion shafts are typically a cantilever-type design that has an enclosed impeller on one end and a tilting-pad bearing on the other. The pinion gear is between these two components. The number of impeller-pinions (i.e., stages) varies with the application and the original equipment vendor. However, all bullgear compressors contain multiple pinions that operate in series. Atmospheric air or gas enters the first-stage pinion, where the pressure is increased by the centrifugal force created by the first-stage impeller. The partially compressed air leaves the first stage, passes through an intercooler, and enters the second-stage impeller. This process is repeated until the fully compressed air leaves through the final pinion-impeller, or stage. Most bullgear compressors are designed to operate with a gear speed of 3,600 rpm. In a typical four-stage compressor, the pinions operate at progressively higher speeds. A typical range is between 12,000 rpm (first stage) and 70,000 rpm (fourth stage). Because of their cantilever design and pinion rotating speeds, bullgear compressors are extremely sensitive to variations in demand or downstream pressure changes. Because of this sensitivity, their use should be limited to baseload applications. Bullgear compressors are not designed for, nor will they tolerate, load-following applications. They should not be installed in the same discharge manifold with positive-displacement compressors, especially reciprocating compressors. The standing-wave pulses created by many positive-displacement compressors create enough variation in the discharge manifold to cause potentially serious instability. In addition, the large helical gear used for the bullgear creates an axial oscillation or thrusting that contributes to instability within the compressor. This axial movement is transmitted throughout the machine-train.

Performance The physical laws of thermodynamics, which define their efficiency and system dynamics, govern compressed-air systems and compressors. This section discusses both the first and second laws of thermodynamics, which apply to all compressors and compressed-air systems. Also applying to

138 Compressors

these systems are the Ideal Gas Law and the concepts of pressure and compression.

First Law of Thermodynamics This law states that energy cannot be created or destroyed during a process, such as compression and delivery of air or gas, although it may change from one form of energy to another. In other words, whenever a quantity of one kind of energy disappears, an exactly equivalent total of other kinds of energy must be produced. This is expressed for a steady-flow open system such as a compressor by the following relationship: Net energy added to system as heat + and work

Stored energy of mass entering − system

Stored energy of mass leaving system

=0

Second Law of Thermodynamics The second law of thermodynamics states that energy exists at various levels and is available for use only if it can move from a higher to a lower level. For example, it is impossible for any device to operate in a cycle and produce work while exchanging heat only with bodies at a single fixed temperature. In thermodynamics a measure of the unavailability of energy has been devised and is known as entropy. As a measure of unavailability, entropy increases as a system loses heat, but it remains constant when there is no gain or loss of heat as in an adiabatic process. It is defined by the following differential equation: dQ dS = T where: T = Temperature (Fahrenheit) Q = Heat added (BTU)

Pressure/Volume/Temperature (PVT) Relationship Pressure, temperature, and volume are properties of gases that are completely interrelated. Boyle’s Law and Charles’ Law may be combined into one equation that is referred to as the Ideal Gas Law. This equation is always true for ideal gases and is true for real gases under certain conditions. P1 V1 P2 V2 = T1 T2

Compressors 139

For air at room temperature, the error in this equation is less than 1% for pressures as high as 400 psia. For air at one atmosphere of pressure, the error is less than 1% for temperatures as low as −200◦ F. These error factors will vary for different gases.

Pressure/Compression In a compressor, pressure is generated by pumping quantities of gas into a tank or other pressure vessel. Progressively increasing the amount of gas in the confined or fixed-volume space increases the pressure. The effects of pressure exerted by a confined gas result from the force acting on the container walls. This force is caused by the rapid and repeated bombardment from the enormous number of molecules that are present in a given quantity of gas. Compression occurs when the space is decreased between the molecules. Less volume means that each particle has a shorter distance to travel, thus proportionately more collisions occur in a given span of time, resulting in a higher pressure. Air compressors are designed to generate particular pressures to meet specific application requirements.

Other Performance Indicators The same performance indicators as those for centrifugal pumps or fans govern centrifugal compressors.

Installation Dynamic compressors seldom pose serious foundation problems. Since moments and shaking forces are not generated during compressor operation, there are no variable loads to be supported by the foundation. A foundation or mounting of sufficient area and mass to maintain compressor level and alignment and to assure safe soil loading is all that is required. The units may be supported on structural steel if necessary. The principles defined for centrifugal pumps also apply to centrifugal compressors. It is necessary to install pressure-relief valves on most dynamic compressors to protect them due to restrictions placed on casing pressure, power input, and to keep out of the compressor’s surge range. Always install a valve capable of bypassing the full-load capacity of the compressor between its discharge port and the first isolation valve.

140 Compressors

Operating Methods The acceptable operating envelope for centrifugal compressors is very limited. Therefore, care should be taken to minimize any variation in suction supply, backpressure caused by changes in demand, and frequency of unloading. The operating guidelines provided in the compressor vendor’s O&M manual should be followed to prevent abnormal operating behavior or premature wear or failure of the system. Centrifugal compressors are designed to be baseloaded and may exhibit abnormal behavior or chronic reliability problems when used in a loadfollowing mode of operation. This is especially true of bullgear and cantilever compressors. For example, a one-psig change in discharge pressure may be enough to cause catastrophic failure of a bullgear compressor. Variations in demand or backpressure on a cantilever design can cause the entire rotating element and its shaft to flex. This not only affects the compressor’s efficiency, but also accelerates wear and may lead to premature shaft or rotor failure. All compressor types have moving parts, high noise levels, high pressures, and high-temperature cylinder and discharge-piping surfaces.

Positive Displacement Positive-displacement compressors can be divided into two major classifications: rotary and reciprocating.

Rotary The rotary compressor is adaptable to direct drive by the use of induction motors or multicylinder gasoline or diesel engines. These compressors are compact, relatively inexpensive, and require a minimum of operating attention and maintenance. They occupy a fraction of the space and weight of a reciprocating machine having equivalent capacity.

Configuration Rotary compressors are classified into three general groups: sliding vane, helical lobe, and liquid-seal ring.

Sliding Vane The basic element of the sliding-vane compressor is the cylindrical housing and the rotor assembly. This compressor, which is illustrated in Figure 8.5,

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Housing

Air inlet

Compressed air out Sliding vane

Figure 8.5 Rotary sliding-vane compressor

has longitudinal vanes that slide radially in a slotted rotor mounted eccentrically in a cylinder. The centrifugal force carries the sliding vanes against the cylindrical case with the vanes forming a number of individual longitudinal cells in the eccentric annulus between the case and rotor. The suction port is located where the longitudinal cells are largest. The size of each cell is reduced by the eccentricity of the rotor as the vanes approach the discharge port, thus compressing the gas. Cyclical opening and closing of the inlet and discharge ports occurs by the rotor’s vanes passing over them. The inlet port is normally a wide opening that is designed to admit gas in the pocket between two vanes. The port closes momentarily when the second vane of each air-containing pocket passes over the inlet port. When running at design pressure, the theoretical operation curves are identical (see Figure 8.6) to those of a reciprocating compressor. However, there is one major difference between a sliding-vane and a reciprocating compressor. The reciprocating unit has spring-loaded valves that open automatically with small pressure differentials between the outside and inside cylinder. The sliding-vane compressor has no valves. The fundamental design considerations of a sliding-vane compressor are the rotor assembly, cylinder housing, and the lubrication system. Housing and Rotor Assembly Cast iron is the standard material used to construct the cylindrical housing, but other materials may be used if corrosive conditions exist. The rotor is usually a continuous piece of steel that includes the shaft and is made from bar stock. Special materials can be selected for corrosive applications. Occasionally, the rotor may be a separate iron casting keyed to a shaft. On most standard air compressors, the rotor-shaft seals are semimetallic packing in a stuffing box. Commercial mechanical rotary seals can be supplied

Pressure

142 Compressors

Design pressure (discharge)

Operation at design pressure

Pressure

Pressure

Volume Discharge pressure Design pressure

Volume Discharge pressure Design pressure

Operation above design pressure

Operation below design pressure

Volume

Figure 8.6 Theoretical operation curves for rotary compressors with built-in porting

when needed. Cylindrical roller bearings are generally used in these assemblies. Vanes are usually asbestos or cotton cloth impregnated with a phenolic resin. Bronze or aluminum also may be used for vane construction. Each vane fits into a milled slot extending the full length of the rotor and slides radially in and out of this slot once per revolution. Vanes are the most maintenanceprone part in the compressor. There are from 8 to 20 vanes on each rotor, depending upon its diameter. A greater number of vanes increase compartmentalization, which reduces the pressure differential across each vane. Lubrication System A V-belt-driven, force-fed oil lubrication system is used on water-cooled compressors. Oil goes to both bearings and to several points in the cylinder. Ten times as much oil is recommended to lubricate the rotary cylinder as is required for the cylinder of a corresponding reciprocating compressor. The oil carried over with the gas to the line may be reduced 50% with an oil separator on the discharge. Use of an aftercooler ahead of the separator permits removal of 85 to 90% of the entrained oil.

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Figure 8.7 Helical lobe, or screw, rotary air compressor

Helical Lobe or Screw The helical lobe, or screw, compressor is shown in Figure 8.7. It has two or more mating sets of lobe-type rotors mounted in a common housing. The male lobe, or rotor, is usually direct-driven by an electric motor. The female lobe, or mating rotor, is driven by a helical gear set that is mounted on the outboard end of the rotor shafts. The gears provide both motive power for the female rotor and absolute timing between the rotors. The rotor set has extremely close mating clearance (i.e., about 0.5 mils) but no metal-to-metal contact. Most of these compressors are designed for oil-free operation. In other words, no oil is used to lubricate or seal the rotors. Instead, oil lubrication is limited to the timing gears and bearings that are outside the air chamber. Because of this, maintaining proper clearance between the two rotors is critical. This type of compressor is classified as a constant volume, variablepressure machine that is quite similar to the vane-type rotary in general characteristics. Both have a built-in compression ratio. Helical-lobe compressors are best suited for base-load applications where they can provide a constant volume and pressure of discharge gas. The only recommended method of volume control is the use of variable-speed motors. With variable-speed drives, capacity variations can be obtained with

144 Compressors

a proportionate reduction in speed. A 50% speed reduction is the maximum permissible control range. Helical-lobe compressors are not designed for frequent or constant cycles between load and no-load operation. Each time the compressor unloads, the rotors tend to thrust axially. Even though the rotors have a substantial thrust bearing and, in some cases, a balancing piston to counteract axial thrust, the axial clearance increases each time the compressor unloads. Over time, this clearance will increase enough to permit a dramatic rise in the impact energy created by axial thrust during the transient from loaded to unloaded conditions. In extreme cases, the energy can be enough to physically push the rotor assembly through the compressor housing. Compression ratio and maximum inlet temperature determine the maximum discharge temperature of these compressors. Discharge temperatures must be limited to prevent excessive distortion between the inlet and discharge ends of the casing and rotor expansion. High-pressure units are water-jacketed in order to obtain uniform casing temperature. Rotors also may be cooled to permit a higher operating temperature. If either casing distortion or rotor expansion occur, the clearance between the rotating parts will decrease, and metal-to-metal contact will occur. Since the rotors typically rotate at speeds between 3,600 and 10,000 rpm, metalto-metal contact normally results in instantaneous, catastrophic compressor failure. Changes in differential pressures can be caused by variations in either inlet or discharge conditions (i.e., temperature, volume, or pressure). Such changes can cause the rotors to become unstable and change the load zones in the shaft-support bearings. The result is premature wear and/or failure of the bearings. Always install a relief valve that is capable of bypassing the full-load capacity of the compressor between its discharge port and the first isolation valve. Since helical-lobe compressors are less tolerant to over-pressure operation, safety valves are usually set within 10% of absolute discharge pressure, or 5 psi, whichever is lower.

Liquid-Seal Ring The liquid-ring, or liquid-piston, compressor is shown in Figure 8.8. It has a rotor with multiple forward-turned blades that rotate about a central cone that contains inlet and discharge ports. Liquid is trapped between adjacent

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Inlet

Discharge

Discharge port

Inlet port Discharge port

Inlet port

Rotation

Figure 8.8 Liquid-seal ring rotary air compressor

blades, which drive the liquid around the inside of an elliptical casing. As the rotor turns, the liquid face moves in and out of this space due to the casing shape, creating a liquid piston. Porting in the central cone is built-in and fixed, and there are no valves. Compression occurs within the pockets or chambers between the blades before the discharge port is uncovered. Since the port location must be designed and built for a specific compression ratio, it tends to operate above or below the design pressure (refer back to Figure 8.6). Liquid-ring compressors are cooled directly rather than by jacketed casing walls. The cooling liquid is fed into the casing where it comes into direct contact with the gas being compressed. The excess liquid is discharged with the gas. The discharged mixture is passed through a conventional baffle or centrifugal-type separator to remove the free liquid. Because of the intimate contact of gas and liquid, the final discharge temperature can be held close to the inlet cooling water temperature. However, the discharge gas is saturated with liquid at the discharge temperature of the liquid. The amount of liquid passed through the compressor is not critical and can be varied to obtain the desired results. The unit will not be damaged if a large quantity of liquid inadvertently enters its suction port. Lubrication is required only in the bearings, which are generally located external to the casing. The liquid itself acts as a lubricant, sealing medium, and coolant for the stuffing boxes.

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Performance Performance of a rotary positive-displacement compressor can be evaluated using the same criteria as a positive-displacement pump. As in constantvolume machines, performance is determined by rotation speed, internal slip, and total backpressure on the compressor. The volumetric output of rotary positive-displacement compressors can be controlled by speed changes. The more slowly the compressor turns, the lower its output volume. This feature permits the use of these compressors in load-following applications. However, care must be taken to prevent sudden, radical changes in speed. Internal slip is simply the amount of gas that can flow through internal clearances from the discharge back to the inlet. Obviously, internal wear will increase internal slip. Discharge pressure is relatively constant regardless of operating speed. With the exceptions of slight pressure variations caused by atmospheric changes and backpressure, a rotary positive-displacement compressor will provide a fixed discharge pressure. Backpressure, which is caused by restrictions in the discharge piping or demand from users of the compressed air or gas, can have a serious impact on compressor performance. If backpressure is too low or demand too high, the compressor will be unable to provide sufficient volume or pressure to the downstream systems. In this instance, the discharge pressure will be noticeably lower than designed. If the backpressure is too high or demand too low, the compressor will generate a discharge pressure higher than designed. It will continue to compress the air or gas until it reaches the unload setting on the system’s relief valve or until the brake horsepower required exceeds the maximum horsepower rating of the driver.

Installation Installation requirements for rotary positive-displacement compressors are similar to those for any rotating machine. Review the installation requirements for centrifugal pumps and compressors for foundation, pressure-relief, and other requirements. As with centrifugal compressors, rotary positive-displacement compressors must be fitted with pressure-relief devices to limit the discharge or interstage pressures to a safe maximum for the equipment served.

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In applications where demand varies, rotary positive-displacement compressors require a downstream receiver tank or reservoir that minimizes the load-unload cycling frequency of the compressor. The receiver tank should have sufficient volume to permit acceptable unload frequencies for the compressor. Refer to the vendor’s O&M manual for specific receiver-tank recommendations.

Operating Methods All compressor types have moving parts, high noise levels, high pressures, and high-temperature cylinder and discharge-piping surfaces. Rotary positive-displacement compressors should be operated as baseloaded units. They are especially sensitive to the repeated start-stop operation required by load-following applications. Generally, rotary positivedisplacement compressors are designed to unload about every six to eight hours. This unload cycle is needed to dissipate the heat generated by the compression process. If the unload frequency is too great, these compressors have a high probability of failure. There are several primary operating control inputs for rotary positivedisplacement compressors. These control inputs are: discharge pressure, pressure fluctuations, and unloading frequency.

Discharge Pressure This type of compressor will continue to compress the air volume in the downstream system until: (1) some component in the system fails; (2) the brake horsepower exceeds the driver’s capacity; or (3) a safety valve opens. Therefore, the operator’s primary control input should be the compressor’s discharge pressure. If the discharge pressure is below the design point, it is a clear indicator that the total downstream demand is greater than the unit’s capacity. If the discharge pressure is too high, the demand is too low, and excessive unloading will be required to prevent failure.

Pressure Fluctuations Fluctuations in the inlet and discharge pressures indicate potential system problems that may adversely affect performance and reliability. Pressure fluctuations are generally caused by changes in the ambient environment, turbulent flow, or restrictions caused by partially blocked inlet filters. Any of these problems will result in performance and reliability problems if not corrected.

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Unloading Frequency The unloading function in rotary positive-displacement compressors is automatic and not under operator control. Generally, a set of limit switches, one monitoring internal temperature and one monitoring discharge pressure, are used to trigger the unload process. By design, the limit switch that monitors the compressor’s internal temperature is the primary control. The secondary control, or discharge-pressure switch, is a fail-safe design to prevent overloading of the compressor. Depending on design, rotary positive-displacement compressors have an internal mechanism designed to minimize the axial thrust caused by the instantaneous change from fully loaded to unloaded operating conditions. In some designs, a balancing piston is used to absorb the rotor’s thrust during this transient. In others, oversized thrust bearings are used. Regardless of the mechanism used, none provides complete protection from the damage imparted by the transition from load to no-load conditions. However, as long as the unload frequency is within design limits, this damage will not adversely affect the compressor’s useful operating life or reliability. However, an unload frequency greater than that accommodated in the design will reduce the useful life of the compressor and may lead to premature, catastrophic failure. Operating practices should minimize, as much as possible, the unload frequency of these compressors. Installation of a receiver tank and modification of user-demand practices are the most effective solutions to this type of problem.

Reciprocating Reciprocating compressors are widely used by industry and are offered in a wide range of sizes and types. They vary from units requiring less than 1 hp to more than 12,000 hp. Pressure capabilities range from low vacuums at intake to special compressors capable of 60,000 psig or higher. Reciprocating compressors are classified as constant-volume, variablepressure machines. They are the most efficient type of compressor and can be used for partial-load, or reduced-capacity, applications. Because of the reciprocating pistons and unbalanced rotating parts, the unit tends to shake. Therefore, it is necessary to provide a mounting that

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stabilizes the installation. The extent of this requirement depends on the type and size of the compressor. Because reciprocating compressors should be supplied with clean gas, inlet filters are recommended in all applications. They cannot satisfactorily handle liquids entrained in the gas, although vapors are no problem if condensation within the cylinders does not take place. Liquids will destroy the lubrication and cause excessive wear. Reciprocating compressors deliver a pulsating flow of gas that can damage downstream equipment or machinery. This is sometimes a disadvantage, but pulsation dampers can be used to alleviate the problem.

Configuration Certain design fundamentals should be clearly understood before analyzing the operating condition of reciprocating compressors. These fundamentals include frame and running gear, inlet and discharge valves, cylinder cooling, and cylinder orientation.

Frame and Running Gear Two basic factors guide frame and running gear design. The first factor is the maximum horsepower to be transmitted through the shaft and running gear to the cylinder pistons. The second factor is the load imposed on the frame parts by the pressure differential between the two sides of each piston. This is often called pin load because this full force is directly exerted on the crosshead and crankpin. These two factors determine the size of bearings, connecting rods, frame, and bolts that must be used throughout the compressor and its support structure.

Cylinder Design Compression efficiency depends entirely upon the design of the cylinder and its valves. Unless the valve area is sufficient to allow gas to enter and leave the cylinder without undue restriction, efficiency cannot be high. Valve placement for free flow of the gas in and out of the cylinder is also important. Both efficiency and maintenance are influenced by the degree of cooling during compression. The method of cylinder cooling must be consistent with the service intended. The cylinders and all the parts must be designed to withstand the maximum application pressure. The most economical materials that will give the proper strength and the longest service under the design conditions are generally used.

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Inlet and Discharge Valves Compressor valves are placed in each cylinder to permit one-way flow of gas, either into or out of the cylinder. There must be one or more valve(s) for inlet and discharge in each compression chamber. Each valve opens and closes once for each revolution of the crankshaft. The valves in a compressor operating at 700 rpm for 8 hours per day and 250 days per year will have cycled (i.e., opened and closed) 42,000 times per hour, 336,000 times per day, or 84 million times in a year. The valves 1 of a second to open, let the gas pass through, and to close. have less than 10 They must cycle with a minimum of resistance for minimum power consumption. However, the valves must have minimal clearance to prevent excessive expansion and reduced volumetric efficiency. They must be tight under extreme pressure and temperature conditions. Finally, the valves must be durable under many kinds of abuse. There are four basic valve designs used in these compressors: finger, channel, leaf, and annular ring. Within each class there may be variations in design, depending upon operating speed and size of valve required. Finger Figure 8.9 is an exploded view of a typical finger valve. These valves are used for smaller, air-cooled compressors. One end of the finger is fixed and the opposite end lifts when the valve opens.

Head Discharge valve Valve plate Inlet valve

Figure 8.9 Finger valve configuration

Cylinder

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Valve closed: A tight seat is formed without slamming or friction, so seat wear is at a minimum. Both channel and spring are precision made to assure a perfect fit. A gas space is formed between the bowed spring and the flat channel.

Valve opening: Channel lifts straight up in the guides without flexing. Opening is even over the full length of the port, giving uniform air velocity without turbulance. Cushioning is effected by the compression and escape of the gas between spring and channel.

Valve wide open: Gas trapped between spring and channel has been compressed and in escaping has allowed channel to float in its stop.

Figure 8.10 Channel valve configuration

Channel The channel valve shown in Figure 8.10 is widely used in mid- to large-sized compressors. This valve uses a series of separate stainless steel channels. As explained in the figure, this is a cushioned valve, which adds greatly to its life. Leaf The leaf valve (see Figure 8.11) has a configuration somewhat like the channel valve. It is made of flat-strip steel that opens against an arched stop plate. This results in valve flexing only at its center with maximum lift. The valve operates as its own spring. Annular Ring Figure 8.12 shows exploded views of typical inlet and discharge annular-ring valves. The valves shown have a single ring, but larger sizes may have two or three rings. In some designs, the concentric rings are tied into a single piece by bridges.

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Figure 8.11 Leaf spring configuration The springs and the valve move into a recess in the stop plate as the valve opens. Gas that is trapped in the recess acts as a cushion and prevents slamming. This eliminates a major source of valve and spring breakage. The valve shown was the first cushioned valve built.

Cylinder Cooling Cylinder heat is produced by the work of compression plus friction, which is caused by the action of the piston and piston rings on the cylinder wall and packing on the rod. The amount of heat generated can be considerable,

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Inlet

Discharge

Figure 8.12 Annular-ring valves

particularly when moderate to high compression ratios are involved. This can result in undesirably high operating temperatures. Most compressors use some method to dissipate a portion of this heat to reduce the cylinder wall and discharge gas temperatures. The following are advantages of cylinder cooling: ●

Lowering cylinder wall and cylinder head temperatures reduces loss of capacity and horsepower per unit volume due to suction gas preheating during inlet stroke. This results in more gas in the cylinder for compression.

●

Reducing cylinder wall and cylinder head temperatures removes more heat from the gas during compression, lowering its final temperature and reducing the power required.

154 Compressors

●

Reducing the gas temperature and that of the metal surrounding the valves results in longer valve service life and reduces the possibility of deposit formation.

●

Reduced cylinder wall temperature promotes better lubrication, resulting in longer life and reduced maintenance.

●

Cooling, particularly water-cooling, maintains a more even temperature around the cylinder bore and reduces warpage.

Cylinder Orientation Orientation of the cylinders in a multistage or multicylinder compressor directly affects the operating dynamics and vibration level. Figure 8.13 illustrates a typical three-piston, air-cooled compressor. Since three pistons are oriented within a 120-degree arc, this type of compressor generates higher vibration levels than the opposed piston compressor illustrated in Figure 8.14.

Suction valve (discharge valve on opposite side) Piston

2nd stage

Discharge valve

Suction valve Air inlet 1st stage

1st stage

Crankcase oil dipstick

Connecting rods

Oil sump Crankshaft

Figure 8.13 Three-piston compressor generates higher vibration levels

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Figure 8.14 Opposed-piston compressor balances piston forces

Performance Reciprocating-compressor performance is governed almost exclusively by operating speed. Each cylinder of the compressor will discharge the same volume, excluding slight variations caused by atmospheric changes, at the same discharge pressure each time it completes the discharge stroke. As the rotation speed of the compressor changes, so does the discharge volume. The only other variables that affect performance are the inlet-discharge valves, which control flow into and out of each cylinder. Although reciprocating compressors can use a variety of valve designs, it is crucial that the valves perform reliably. If they are damaged and fail to operate at the

156 Compressors

proper time or do not seal properly, overall compressor performance will be substantially reduced.

Installation A carefully planned and executed installation is extremely important and makes compressor operation and maintenance easier and safer. Key components of a compressor installation are location, foundation, and piping.

Location The preferred location for any compressor is near the center of its load. However, the choice is often influenced by the cost of supervision, which can vary by location. The ongoing cost of supervision may be less expensive at a less-optimum location, which can offset the cost of longer piping. A compressor will always give better, more reliable service when enclosed in a building that protects it from cold, dusty, damp, and corrosive conditions. In certain locations it may be economical to use a roof only, but this is not recommended unless the weather is extremely mild. Even then, it is crucial to prevent rain and wind-blown debris from entering the moving parts. Subjecting a compressor to adverse inlet conditions will dramatically reduce reliability and significantly increase maintenance requirements. Ventilation around a compressor is vital. On a motor-driven, air-cooled unit, the heat radiated to the surrounding air is at least 65% of the power input. On a water-jacketed unit with an aftercooler and outside receiver, the heat radiated to the surrounding air may be 15 to 25% of the total energy input, which is still a substantial amount of heat. Positive outside ventilation is recommended for any compressor room where the ambient temperature may exceed 104◦ F.

Foundation Because of the alternating movement of pistons and other components, reciprocating compressors often develop a shaking that alternates in direction. This force must be damped and contained by the mounting. The foundation also must support the weight load of the compressor and its driver. There are many compressor arrangements, and the net magnitude of the moments and forces developed can vary a great deal among them. In some cases, they are partially or completely balanced within the compressors themselves. In others, the foundation must handle much of the force.

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When complete balance is possible, reciprocating compressors can be mounted on a foundation just large and rigid enough to carry the weight and maintain alignment. However, most reciprocating compressors require larger, more massive foundations than other machinery. Depending upon size and type of unit, the mounting may vary from simply bolting to the floor to attaching to a massive foundation designed specifically for the application. A proper foundation must: (1) maintain the alignment and level of the compressor and its driver at the proper elevation, and (2) minimize vibration and prevent its transmission to adjacent building structures and machinery. There are five steps to accomplish the first objective: 1

The safe weight-bearing capacity of the soil must not be exceeded at any point on the foundation base.

2

The load to the soil must be distributed over the entire area.

3

The size and proportion of the foundation block must be such that the resultant vertical load due to the compressor, block, and any unbalanced force falls within the base area.

4

The foundation must have sufficient mass and weight-bearing area to prevent its sliding on the soil due to unbalanced forces.

5

Foundation temperature must be uniform to prevent warping.

Bulk is not usually the complete solution to foundation problems. A certain weight is sometimes necessary, but soil area is usually of more value than foundation mass. Determining if two or more compressors should have separate or single foundations depends on the compressor type. A combined foundation is recommended for reciprocating units since the forces from one unit usually will partially balance out the forces from the others. In addition, the greater mass and surface area in contact with the ground damps foundation movement and provides greater stability. Soil quality may vary seasonally, and such conditions must be carefully considered in the foundation design. No foundation should rest partially on bedrock and partially on soil; it should rest entirely on one or the other. If placed on the ground, make sure that part of the foundation does not rest on soil that has been disturbed. In addition, pilings may be necessary to ensure stability.

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Piping Piping should easily fit the compressor connections without needing to spring or twist it to fit. It must be supported independently of the compressor and anchored, as necessary, to limit vibration and to prevent expansion strains. Improperly installed piping may distort or pull the compressor’s cylinders or casing out of alignment.

Air Inlet The intake pipe on an air compressor should be as short and direct as possible. If the total run of the inlet piping is unavoidably long, the diameter should be increased. The pipe size should be greater than the compressor’s air-inlet connection. Cool inlet air is desirable. For every 5◦ F of ambient air temperature reduction, the volume of compressed air generated increases by 1% with the same power consumption. This increase in performance is due to the greater density of the intake air. It is preferable for the intake air to be taken from outdoors. This reduces heating and air conditioning costs and, if properly designed, has fewer contaminants. However, the intake piping should be a minimum of six feet above the ground and be screened or, preferably, filtered. An air inlet must be free of steam and engine exhausts. The inlet should be hooded or turned down to prevent the entry of rain or snow. It should be above the building eaves and several feet from the building.

Discharge Discharge piping should be the full size of the compressor’s discharge connection. The pipe size should not be reduced until the point along the pipeline is reached where the flow has become steady and nonpulsating. With a reciprocating compressor, this is generally beyond the aftercooler or the receiver. Pipes to handle nonpulsating flow are sized by normal methods, and long-radius bends are recommended. All discharge piping must be designed to allow adequate expansion loops or bends to prevent undue stresses at the compressor.

Drainage Before piping is installed, the layout should be analyzed to eliminate low points where liquid could collect and to provide drains where low points cannot be eliminated. A regular part of the operating procedure must be

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the periodic drainage of low points in the piping and separators, as well as inspection of automatic drain traps.

Pressure-Relief Valves All reciprocating compressors must be fitted with pressure relief devices to limit the discharge or interstage pressures to a safe maximum for the equipment served. Always install a relief valve that is capable of bypassing the full-load capacity of the compressor between its discharge port and the first isolation valve. The safety valves should be set to open at a pressure slightly higher than the normal discharge-pressure rating of the compressor. For standard 100- to 115-psig two-stage air compressors, safety valves are normally set at 125 psig. The pressure-relief safety valve is normally situated on top of the air reservoir, and there must be no restriction on its operation. The valve is usually of the “huddling chamber” design, in which the static pressure acting on its disk area causes it to open. Figure 8.15 illustrates how such a valve functions. As the valve pops, the air space within the huddling chamber between

3. Spring holds piston closed.

4. When the valve setting is reached, the poppet “opens” limiting pressure in upper chamber.

7. Vent connection permits unloading pump through relief valve.

1. Inlet pressure here...

2. Is sensed above piston and at pilot valve through orifice in piston. View A closed

6. Piston moves up to divert pump output directly to tank. View B cracked

Figure 8.15 Illustrates how a safety valve functions

5. When this pressure is 20 psi higher than in upper chamber ... View C Relieving

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the seat and blowdown ring fills with pressurized air and builds up more pressure on the roof of the disk holder. This temporary pressure increases the upward thrust against the spring, causing the disk and its holder to fully pop open. Once a predetermined pressure drop (i.e., blowdown) occurs, the valve closes with a positive action by trapping pressurized air on top of the disk holder. Raising or lowering the blowdown ring adjusts the pressuredrop setpoint. Raising the ring increases the pressure-drop setting, while lowering it decreases the setting.

Operating Methods Compressors can be hazardous to work around because they have moving parts. Ensure that clothing is kept away from belt drives, couplings, and exposed shafts. In addition, high-temperature surfaces around cylinders and discharge piping are exposed. Compressors are notoriously noisy, so ear protection should be worn. These machines are used to generate highpressure gas so, when working around them, it is important to wear safety glasses and to avoid searching for leaks with bare hands. High-pressure leaks can cause severe friction burns.

Troubleshooting Compressors can be divided into three classifications: centrifugal, rotary, and reciprocating. This section identifies the common failure modes for each.

Centrifugal The operating dynamics of centrifugal compressors are the same as for other centrifugal machine-trains. The dominant forces and vibration profiles are typically identical to pumps or fans. However, the effects of variable load and other process variables (e.g., temperatures, inlet/discharge pressure, etc.) are more pronounced than in other rotating machines. Table 8.1 identifies the common failure modes for centrifugal compressors. Aerodynamic instability is the most common failure mode for centrifugal compressors. Variable demand and restrictions of the inlet-air flow are common sources of this instability. Even slight variations can cause dramatic changes in the operating stability of the compressor.

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Table 8.1 Common failure modes of centrifugal compressors

•

•

Motor trips

Water in lube oil

Persistent unloading

• •

Change in system resistance

• •

Clogged oil strainer/filter

•

Compressor not up to speed

•

Condensate in oil reservoir Damaged rotor

•

Dry gear coupling

•

Excessive bearing clearance

•

Excessive inlet temperature

Units do not stay in alignment

• •

Build-up of deposits on diffuser Build-up of deposits on rotor

Excessive bearing oil drain temp.

•

Bearing lube oil orifice missing or plugged Bent rotor (caused by uneven heating and cooling)

Low lube oil pressure

Loss of discharge pressure

Compressor surges

THE CAUSES

Excessive vibration

THE PROBLEM

•

Failure of both main and auxiliary oil pumps

•

Faulty temperature gauge or switch

•

•

• Continued

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Table 8.1 continued

Improperly assembled parts

•

•

• •

Insufficient flow

•

Leak in discharge piping

•

Leak in lube oil cooler tubes or tube sheet

•

Leak in oil pump suction piping Liquid “slugging”

•

Loose or broken bolting

•

Loose rotor parts

•

•

Oil leakage

•

Oil pump suction plugged

•

Oil reservoir low level

•

Operating at low speed w/o auxiliary oil pump

• •

Motor trips

Water in lube oil

•

Incorrect pressure control valve setting

Operating in critical speed range






Compressor surges

THE CAUSES

Excessive vibration

THE PROBLEM

Compressors 163

Table 8.1 continued


Water in lube oil

Motor trips

•

•

•

•

Relief valve improperly set or stuck open

•

• •

Rough rotor shaft journal surface Shaft misalignment

•

Sympathetic vibration

•

•

•

• •

•

•

Vibration

•

Warped foundation or baseplate

•

Wiped or damaged bearings Worn or damaged coupling

• •

Poor oil condition

Rotor imbalance


•


Piping strain


•


Operating in surge region

Compressor surges

THE CAUSES

Excessive vibration

THE PROBLEM

• •

•

Entrained liquids and solids also can affect operating life. When dirty air must be handled, open-type impellers should be used. An open design provides the ability to handle a moderate amount of dirt or other solids in the inlet-air supply. However, inlet filters are recommended for all

164 Compressors

applications, and controlled liquid injection for cleaning and cooling should be considered during the design process.

Rotary-Type, Positive Displacement Table 8.2 lists the common failure modes of rotary-type, positivedisplacement compressors. This type of compressor can be grouped into two types: sliding vane and rotary screw.

Sliding Vane Compressors Sliding-vane compressors have the same failure modes as vane-type pumps. The dominant components in their vibration profile are running speed, vane-pass frequency, and bearing-rotation frequencies. In normal operation, the dominant energy is at the shaft’s running speed. The other frequency components are at much lower energy levels. Common failures of this type of compressor occur with shaft seals, vanes, and bearings. Shaft Seals Leakage through the shaft’s seals should be checked visually once a week or as part of every data-acquisition route. Leakage may not be apparent from the outside of the gland. If the fluid is removed through a vent, the discharge should be configured for easy inspection. Generally, more leakage than normal is the signal to replace a seal. Under good conditions, they have a normal life of 10,000 to 15,000 hours and should routinely be replaced when this service life has been reached. Vanes Vanes wear continuously on their outer edges and, to some degree, on the faces that slide in and out of the slots. The vane material is affected somewhat by prolonged heat, which causes gradual deterioration. Typical life expectancy of vanes in 100-psig services is about 16,000 hours of operation. For low-pressure applications, life may reach 32,000 hours. Replacing vanes before they break is extremely important. Breakage during operation can severely damage the compressor, which requires a complete overhaul and realignment of heads and clearances. Bearings In normal service, bearings have a relatively long life. Replacement after about six years of operation is generally recommended. Bearing defects are usually displayed in the same manner in a vibration profile as for any rotating machine-train. Inner and outer race defects are the dominant failure modes, but roller spin also may contribute to the failure.

Compressors 165

Table 8.2 Common failure modes of rotary-type, positive-displacement compressors

Excessive discharge pressure

•

Excessive inlet temperature/moisture

•

Insufficient suction air/gas supply Internal component wear

•

Motor or driver failure

•

•

•

•

•

•

•

•

•

•

•

•

•

•

•

•

Rotating element binding

•

•

•

•

Solids or dirt in inlet air/gas supply

•

•

Relief valve stuck open or set wrong

Speed too low

• •

•

•

Wrong direction of rotation

•

•

Elevated air/gas temperature

Elevated motor temperature

•

•

•

•

•

Suction filter or strainer clogged

•

•

•

Pipe strain on compressor casing

Motor trips

Excessive power demand

• •

Coupling misaligned

Excessive vibration and noise

•

Excessive heat

•

Excessive wear

Insufficient capacity

Air leakage into suction piping or shaft seal

Insufficient discharge pressure

THE CAUSES

No air/gas delivery

THE PROBLEM

•

•

• •

• •

166 Compressors

Rotary Screw The most common reason for compressor failure or component damage is process instability. Rotary-screw compressors are designed to deliver a constant volume and pressure of air or gas. These units are extremely susceptible to any change in either inlet or discharge conditions. A slight variation in pressure, temperature, or volume can result in instantaneous failure. The following are used as indices of instability and potential problems: rotor mesh, axial movement, thrust bearings, and gear mesh.

Rotor Mesh In normal operation, the vibration energy generated by male and female rotor meshing is very low. As the process becomes unstable, the energy due to the rotor-meshing frequency increases, with both the amplitude of the meshing frequency and the width of the peak increasing. In addition, the noise floor surrounding the meshing frequency becomes more pronounced. This white noise is similar to that observed in a cavitating pump or unstable fan.

Axial Movement The normal tendency of the rotors and helical timing gears is to generate axial shaft movement, or thrusting. However, the extremely tight clearances between the male and female rotors do not tolerate any excessive axial movement, and therefore, axial movement should be a primary monitoring parameter. Axial measurements are needed from both rotor assemblies. If there is any increase in the vibration amplitude of these measurements, it is highly probable that the compressor will fail.

Thrust Bearings While process instability can affect both the fixed and float bearings, the thrust bearing is more likely to show early degradation as a result of process instability or abnormal compressor dynamics. Therefore, these bearings should be monitored closely, and any degradation or hint of excessive axial clearance should be corrected immediately.

Gear Mesh The gear mesh vibration profile also provides an indication of prolonged compressor instability. Deflection of the rotor shafts changes the wear pattern on the helical gear sets. This change in pattern increases the backlash in the gear mesh, results in higher vibration levels, and increases thrusting.

Compressors 167

Reciprocating, Positive Displacement Reciprocating compressors have a history of chronic failures that include valves, lubrication system, pulsation, and imbalance. Table 8.3 identifies common failure modes and causes for this type of compressor. Like all reciprocating machines, reciprocating compressors normally generate higher levels of vibration than centrifugal machines. In part, the increased level of vibration is due to the impact as each piston reaches top dead center and bottom dead center of its stroke. The energy levels also are influenced by the unbalanced forces generated by nonopposed pistons and looseness in the piston rods, wrist pins, and journals of the compressor. In most cases, the dominant vibration frequency is the second harmonic (2X) of the main crankshaft’s rotating speed. Again, this results from the impact that occurs when each piston changes directions (i.e., two impacts occur during one complete crankshaft rotation).

Valves Valve failure is the dominant failure mode for reciprocating compressors. Because of their high cyclic rate, which exceeds 80 million cycles per year, inlet and discharge valves tend to work harder and crack.

Lubrication System Poor maintenance of lubrication-system components, such as filters and strainers, typically causes premature failure. Such maintenance is crucial to reciprocating compressors because they rely on the lubrication system to provide a uniform oil film between closely fitting parts (e.g., piston rings and the cylinder wall). Partial or complete failure of the lube system results in catastrophic failure of the compressor.

Pulsation Reciprocating compressors generate pulses of compressed air or gas that are discharged into the piping that transports the air or gas to its point(s) of use. This pulsation often generates resonance in the piping system, and pulse impact (i.e., standing waves) can severely damage other machinery connected to the compressed-air system. While this behavior does not cause the compressor to fail, it must be prevented to protect other plant equipment. Note, however, that most compressed-air systems do not use pulsation dampers.

THE CAUSES

Air flow to fan blocked

Ambient temperature too high

Assembly incorrect

Air discharge temperature too high

•

Air filter defective

•

• •

• •

Air leak into pump suction

•

• •

Valve wear and breakage normal

Starts too often

Receiver safety valve pops

Receiver pressure above normal

Piston rod or packing wear excessive

Piston ring, piston, cylinder wear excessive

Outlet water temperature above normal

Operating cycle abnormally long

Oil pumping excessive (single-acting compressor)

Motor over-heating

Intercooler safety valve pops

Intercooler pressure below normal

Intercooler pressure above normal

Excessive compressor vibration

Discharge pressure below normal

Delivery less than rated capacity

Crankcase water accumulation

Crankcase oil pressure low

Compressor parts overheat

Compressor noisy or knocks

Compressor fails to unload

Compressor fails to start

Carbonaceous deposits abnormal

Air discharge temperature above normal

THE PROBLEM

168 Compressors

Table 8.3 A-E in electronic files

• •

• •

• •

Bearings need adjustment or renewal

• • •

Belts slipping

•

• •

•

Belts too tight

•

•

• •

Centrifugal pilot valve leaks

•

Check or discharge valve defective

•

Control air filter, strainer clogged Control air line clogged

•

Control air pipe leaks

• • •

Crankcase oil pressure too high

•

Crankshaft end play too great Cylinder, head, cooler dirty

• • •

Cylinder, head, intercooler dirty

• •

• •

• • •

•H •L •H •L • •

•H •H

•

Detergent oil being used (3)

•

Demand too steady (2) Dirt, rust entering cylinder

•

•

•

•

Compressors 169

Cylinder (piston) worn or scored

•

Discharge line restricted THE CAUSES

Discharge pressure above rating

Electrical conditions wrong

Excitation inadequate

Foundation bolts loose

• •

•

• • •

•

Excessive number of starts

• •

•

• • • • •

•

•

• • • • • • •


Starts too often








Motor over-heating















THE PROBLEM

170 Compressors

Table 8.3 continued

•

Foundation too small

•

Foundation uneven-unit rocks

•

Fuses blown Gaskets leak

•

• •

• •

• •

Gauge defective

•

Gear pump worn/defective

•

•

Intake pipe restricted, too small, too long

•

•

• •

• •

•

•

• •

•

• •

• •

•

•

• •

• •

Intercooler leaks

•

Intercooler passages clogged

•

• •

Leveling wedges left under compressor

•

•

•

•

•

Compressors 171

•

Intercooler pressure too high

Liquid carry-over

•H •H

•

•

Intercooler, drain more often

Intercooler vibrating

•

•

Grout, improperly placed Intake filter clogged

•H •L •H •L

THE CAUSES

Lubrication inadequate

Motor overload relay tripped

Motor too small

Low oil pressure relay open

•

Motor rotor loose on shaft

• •

Location too humid and damp

• • •

• •

•

• • •


Starts too often








Motor over-heating















THE PROBLEM

172 Compressors

Table 8.3 continued

•

• •

•

New valve on worn seat “Off ” time insufficient Oil feed excessive

• • •

• •

• •

• •

•

• • •

• •

•

•

Oil wrong type

•

Packing rings worn, stuck, broken

•

Piping improperly supported

•

Piston or piston nut loose

•

Piston ring gaps not staggered

• • •

• • •

• • •

•H •L •H •L • •

•H •H

Compressors 173

Piston or ring drain hole clogged

Piston-to-head clearance too small

•

•

Oil relief valve defective

Piston rings worn, broken, or stuck

•

•

• •

Oil level too low

Oil viscosity incorrect

•

•

Oil filter or strainer clogged Oil level too high

•

THE CAUSES

Rod packing leaks

Regulation piping clogged

•

Pulley or flywheel loose

• •

•

Resonant pulsation (inlet or discharge)

• • • • • •

Receiver, drain more often

•

Receiver too small

•


Starts too often








Motor over-heating















THE PROBLEM

174 Compressors

Table 8.3 continued

•

• •

•

Rod packing too tight

•

Rod scored, pitted, worn Rotation wrong

• • • •

Runs too little (2)

• •

Safety valve defective Safety valve leaks

•

•

• •

•

• •

•

Safety valve set too low

• •

Speed demands exceed rating

• •

Speed lower than rating Speed too high

• •

•

•

•

• •

Springs broken

•

•

• •

•

•

•

System leakage excessive

•

•

• •

•

•

•

Tank ringing noise

• •

Unloader running time too long (1) Unloader or control defective

•

• • • • • •

• • • • • • •

•

• • • • • •

Compressors 175

System demand exceeds rating

THE CAUSES

Unloader setting incorrect

V-belt or other misalignment

Unloader parts worn or dirty

• • • • •

• •

Valves dirty

• • •

Valves incorrectly located

• • • • • • •H •L •H •L • •H •H

Valves not seated in cylinder

• • • • • • •H •L •H •L • •H •H

• • • •

• •

Piston rod or packing wear excessive Receiver pressure above normal

• • • • •


Starts too often



•




Motor over-heating

•


• •


•













THE PROBLEM

176 Compressors

Table 8.3 continued

•

Valves worn or broken

• •

• •

Ventilation poor

• •

•

Water jacket or cooler dirty

• •

• •H •H •H

• •

•

•

•

•

•

•

•

•

• • •

Water jackets or intercooler dirty Water quantity insufficient

•H •L •H •L

•

Voltage abnormally low Water inlet temperature too high

• •

•

•

•

•

Wiring incorrect

•

Worn valve on good seat Wrong oil type

•

•

•

(1) Use automatic start/stop control

(3) Change to nondetergent oil H (In high pressure cylinder) L (In low pressure cylinder)

Compressors 177

(2) Use constant speed control

178 Compressors

Crank arrangements Single crank

Forces Primary Secondary F′ without counterwts. F′′ 0.5F′ with counterwts.

Two cranks at 180°

Couples Primary Secondary None F′D without counterwts. F′D/2 with counterwts.

None

In line cylinders

Zero

2F′′

Opposed cylinders

Zero

Zero

Nil

Nil

141F′ without counterwts. 0.707F′ with counterwts.

Zero

707F′D without counterwts. 0.354F′D with counterwts.

F′′D

1.41F′′

Nil

Nil

Zero

None

Nil

Two cranks at 90°

Two cylinders on one crank Cylinders at 90° Two cylinders on one crank Opposed cylinders

F′ without counterwts. Zero with counterwts. 2F′ without counterwts. F′ with counterwts.

Three cranks at 120° Zero

Zero

Zero

4F′′

Zero

Zero

3.46F′D without counterwts. 1.73F′D with counterwts.

None

3.46F′′D

Four cylinders Cranks at 180° Cranks at 90°

Zero 1.41F′D without counterwts.

Zero

4.0F′′D

0.707F′D with counterwts.

Six cylinders Zero

Zero

Zero

Zero

F′ = Primary inertia force in lbs. F′ = .0000284 RN2W F′′ = Secondary inertia force in lbs. F′′ = R/L F′ R = Crank radius, inches N = R.P.M W = Reciprocating weight of one cylinder, lbs L = Length of connecting rod, inches D = Cylinder center distance

Figure 8.16 Unbalanced inertial forces and couples for various reciprocating compressors

Compressors 179

Each time the compressor discharges compressed air, the air tends to act like a compression spring. Because it rapidly expands to fill the discharge piping’s available volume, the pulse of high-pressure air can cause serious damage. The pulsation wavelength, λ, from a compressor having a doubleacting piston design can be determined by: λ=

60a 34, 050 = 2n n

Where: λ = Wavelength, feet a = Speed of sound = 1,135 feet/second n = Compressor speed, revolutions/minute For a double-acting piston design, a compressor running at 1,200 rpm will generate a standing wave of 28.4 feet. In other words, a shock load equivalent to the discharge pressure will be transmitted to any piping or machine connected to the discharge piping and located within twenty-eight feet of the compressor. Note that for a single-acting cylinder, the wavelength will be twice as long.

Imbalance Compressor inertial forces may have two effects on the operating dynamics of a reciprocating compressor, affecting its balance characteristics. The first is a force in the direction of the piston movement, which is displayed as impacts in a vibration profile as the piston reaches top and bottom dead center of its stroke. The second effect is a couple, or moment, caused by an offset between the axes of two or more pistons on a common crankshaft. The interrelationship and magnitude of these two effects depend upon such factors as: (1) number of cranks; (2) longitudinal and angular arrangement; (3) cylinder arrangement; and (4) amount of counterbalancing possible. Two significant vibration periods result, the primary at the compressor’s rotation speed (X) and the secondary at 2X. Although the forces developed are sinusoidal, only the maximum (i.e., the amplitude) is considered in the analysis. Figure 8.16 shows relative values of the inertial forces for various compressor arrangements.

9

Control Valves

Control valves can be broken into two major classifications: process and fluid power. Process valves control the flow of gases and liquids through a process system. Fluid-power valves control pneumatic or hydraulic systems.

Process Process-control valves are available in a variety of sizes, configurations, and materials of construction. Generally, this type of valve is classified by its internal configuration.

Configuration The device used to control flow through a valve varies with its intended function. The more common types are: ball, gate, butterfly, and globe valves.

Ball Ball valves (see Figure 9.1) are simple shutoff devices that use a ball to stop and start the flow of fluid downstream of the valve. As the valve stem turns to the open position, the ball rotates to a point where part or the entire hole machined through the ball is in line with the valve-body inlet and outlet. This allows fluid to pass through the valve. When the ball rotates so that the hole is perpendicular to the flow path, the flow stops. Most ball valves are quick-acting and require a 90-degree turn of the actuator lever to fully open or close the valve. This feature, coupled with the turbulent flow generated when the ball opening is only partially open, limits the use of the ball valve. Use should be limited to strictly an “on-off ” control function (i.e., fully open or fully closed) because of the turbulent-flow condition and severe friction loss when in the partially open position. They should not be used for throttling or flow-control applications. Ball valves used in process applications may incorporate a variety of actuators to provide direct or remote control of the valve. The more common actuators are either manual or motor-operated. Manual values have a handwheel or lever attached directly or through a gearbox to the valve stem.

Control Valves 181

Thrust washer Lever actuator Bonnet Body

Seat

Ball

Figure 9.1 Ball valve

The valve is opened or closed by moving the valve stem through a 90-degree arc. Motor-controlled valves replace the handwheel with a fractional horsepower motor that can be controlled remotely. The motor-operated valve operates in exactly the same way as the manually operated valve.

Gate Gate valves are used when straight-line, laminar fluid flow and minimum restrictions are needed. These valves use a wedge-shaped sliding plate in the valve body to stop, throttle, or permit full flow of fluids through the valve. When the valve is wide open, the gate is completely inside the valve bonnet. This leaves the flow passage through the valve fully open with no flow restrictions, allowing little or no pressure drop through the valve. Gate valves are not suitable for throttling the flow volume unless specifically authorized for this application by the manufacturer. They generally are not suitable because the flow of fluid through a partially open gate can cause extensive damage to the valve. Gate valves are classified as either rising-stem or nonrising-stem. In the nonrising-stem valve, which is shown in Figure 9.2, the stem is threaded into the gate. As the handwheel on the stem is rotated, the gate travels up or down the stem on the threads, while the stem remains vertically stationary.

182 Control Valves

Closed

Figure 9.2 Nonrising-stem gate valve

This type of valve will almost always have a pointer indicator threaded onto the upper end of the stem to indicate the position of the gate. Valves with rising stems (see Figure 9.3) are used when it is important to know by immediate inspection if the valve is open or closed, or when the threads exposed to the fluid could become damaged by fluid contamination. In this valve, the stem rises out of the valve bonnet when the valve is opened.

Butterfly The butterfly valve has a disk-shaped element that rotates about a central shaft or stem. When the valve is closed, the disk face is across the pipe and blocks the flow. Depending upon the type of butterfly valve, the seat may consist of a bonded resilient liner, a mechanically fastened resilient liner, an insert-type reinforced resilient liner, or an integral metal seat with an O-ring inserted around the edge of the disk. As shown in Figure 9.4, both the full open and the throttled positions permit almost unrestricted flow. Therefore, this valve does not induce turbulent flow in the partially closed position. While the design does not permit exact

Control Valves 183

Figure 9.3 Rising stem gate valve

flow-control capabilities, a butterfly valve can be used for throttling flow through the valve. In addition, these valves have the lowest pressure drop of all the conventional types. For these reasons, they are commonly used in process-control applications.

Globe The globe valve gets its name from the shape of the valve body, although other types of valves also may have globular-shaped bodies. Figure 9.5 shows three configurations of this type of valve: straight-flow, angle-flow, and crossflow. A disk attached to the valve stem controls flow in a globe valve. Turning the valve stem until the disk is seated, which is illustrated in View A of Figure 9.6, closes the valve. The edge of the disk and the seat are very accurately machined to form a tight seal. It is important for globe valves to be installed with the pressure against the disk face to protect the stem packing from system pressure when the valve is shut. While this type of valve is commonly used in the fully open or fully closed position, it also may be used for throttling. However, since the seating surface is a relatively large area, it is not suitable for throttling applications where fine adjustments are required.

184 Control Valves

Figure 9.4 Butterfly valves provide almost unrestricted flow

Straight-flow

Angle-flow

Cross-flow

Figure 9.5 Three globe valve configurations: straight-flow, angle-flow, and cross-flow

Control Valves 185

View A

View B

Figure 9.6 Globe valve When the valve is open, as illustrated in View B of Figure 9.6, the fluid flows through the space between the edge of the disk and the seat. Since the fluid flow is equal on all sides of the center of support when the valve is open, there is no unbalanced pressure on the disk to cause uneven wear. The rate at which fluid flows through the valve is regulated by the position of the disk in relation to the valve seat. The globe valve should never be jammed in the open position. After a valve is fully opened, the handwheel or actuating handle should be closed approximately one-half turn. If this is not done, the valve may seize in the open position making it difficult, if not impossible, to close the valve. Many valves are damaged in the manner. Another reason to partially close a globe valve is because it can be difficult to tell if the valve is open or closed. If jammed in the open position, the stem can be damaged or broken by someone who thinks the valve is closed.

Performance Process-control valves have few measurable criteria that can be used to determine their performance. Obviously, the valve must provide a positive seal when closed. In addition, it must provide a relatively laminar flow with minimum pressure drop in the fully open position. When evaluating valves, the following criteria should be considered: capacity rating, flow characteristics, pressure drop, and response characteristics.

Capacity Rating The primary selection criterion of a control valve is its capacity rating. Each type of valve is available in a variety of sizes to handle most typical

186 Control Valves

process-flow rates. However, proper size selection is critical to the performance characteristics of the valve and the system where it is installed. A valve’s capacity must accommodate variations in viscosity, temperature, flow rates, and upstream pressure.

Flow Characteristics The internal design of process-control valves has a direct impact on the flow characteristics of the gas or liquid flowing through the valve. A fully open butterfly or gate valve provides a relatively straight, obstruction-free flow path. As a result, the product should not be affected.

Pressure Drop Control-valve configuration impacts the resistance to flow through the valve. The amount of resistance, or pressure drop, will vary greatly, depending on type, size, and position of the valve’s flow-control device (i.e. ball, gate, disk). Pressure-drop formulas can be obtained for all common valve types from several sources (e.g., Crane, Technical Paper No. 410).

Response Characteristics With the exception of simple, manually controlled shutoff valves, processcontrol valves are generally used to control the volume and pressure of gases or liquids within a process system. In most applications, valves are controlled from a remote location through the use of pneumatic, hydraulic, or electronic actuators. Actuators are used to position the gate, ball, or disk that starts, stops, directs, or proportions the flow of gas or liquid through the valve. Therefore, the response characteristics of a valve are determined, in part, by the actuator. Three factors critical to proper valve operation are: response time, length of travel, and repeatability. Response Time Response time is the total time required for a valve to open or close to a specific set-point position. These positions are fully open, fully closed, and any position in between. The selection and maintenance of the actuator used to control process-control valves have a major impact on response time. Length of Travel The valve’s flow-control device (i.e., gate, ball, or disk) must travel some distance when going from one set point to another. With a manually operated valve, this is a relatively simple operation. The operator moves

Control Valves 187

the stem lever or handwheel until the desired position is reached. The only reasons a manually controlled valve will not position properly are mechanical wear or looseness between the lever or handwheel and the disk, ball, or gate. For remotely controlled valves, however, there are other variables that directly impact valve travel. These variables depend on the type of actuator that is used. There are three major types of actuators: pneumatic, hydraulic, and electronic. Pneumatic Actuators Pneumatic actuators, including diaphragms, air motors, and cylinders, are suitable for simple on-off valve applications. As long as there is enough air volume and pressure to activate the actuator, the valve can be repositioned over its full length of travel. However, when the air supply required to power the actuator is inadequate or the process-system pressure is too great, the actuator’s ability to operate the valve properly is severely reduced. A pneumatic (i.e., compressed air-driven) actuator is shown in Figure 9.7. This type is not suited for precision flow-control applications, because the compressibility of air prevents it from providing smooth, accurate valve positioning. Hydraulic Actuators Hydraulic (i.e., fluid-driven) actuators, also illustrated in Figure 9.7, can provide a positive means of controlling process valves in most applications. Properly installed and maintained, this type of actuator can provide

Pneumatic or hydraulic cylinder actuator

Figure 9.7 Pneumatic or hydraulic cylinders are used as actuators

188 Control Valves

Motor actuator

Figure 9.8 High-torque electric motors can be used as actuators

accurate, repeatable positioning of the control valve over its full range of travel. Electronic Actuators Some control valves use high-torque electric motors as their actuator (see Figure 9.8). If the motors are properly sized and their control circuits are maintained, this type of actuator can provide reliable, positive control over the full range of travel. Repeatability Repeatability is, perhaps, the most important performance criterion of a process-control valve. This is especially true in applications where precise flow or pressure control is needed for optimum performance of the process system. New process-control valves generally provide the repeatability required. However, proper maintenance and periodic calibration of the valves and their actuators are required to ensure long-term performance. This is

Control Valves 189

especially true for valves that use mechanical linkages as part of the actuator assembly.

Installation Process-control valves cannot tolerate solids, especially abrasives, in the gas or liquid stream. In applications where high concentrations of particulates are present, valves tend to experience chronic leakage or seal problems because the particulate matter prevents the ball, disk, or gate from completely closing against the stationary surface. Simply installing a valve with the same inlet and discharge size as the piping used in the process is not acceptable. In most cases, the valve must be larger than the piping to compensate for flow restrictions within the valve.

Operating Methods Operating methods for control valves, which are designed to control or direct gas and liquid flow through process systems or fluid-power circuits, range from manual to remote, automatic operation. The key parameters that govern the operation of valves are the speed of the control movement and the impact of speed on the system. This is especially important in process systems. Hydraulic hammer, or the shock wave generated by the rapid change in the flow rate of liquids within a pipe or vessel, has a serious and negative impact on all components of the process system. For example, instantaneously closing a large flow-control valve may generate in excess of three million foot-pounds of force on the entire system upstream of the valve. This shock wave can cause catastrophic failure of upstream valves, pumps, welds, and other system components. Changes in flow rate, pressure, direction, and other controllable variables must be gradual enough to permit a smooth transition. Abrupt changes in valve position should be avoided. Neither the valve installation nor the control mechanism should permit complete shutoff, referred to as deadheading, of any circuit in a process system. Restricted flow forces system components, such as pumps, to operate outside of their performance envelope. This reduces equipment reliability and sets the stage for catastrophic failure or abnormal system performance. In applications where radical changes in flow are required for normal system operation, control valves should be configured to provide an adequate bypass for surplus flow in order to protect the system.

190 Control Valves

Poppet

Spring

No flow Body

In

Out

Free flow

Figure 9.9 One-way, fluid-power valve For example, systems that must have close control of flow should use two proportioning valves that act in tandem to maintain a balanced hydraulic or aerodynamic system. The primary, or master, valve should control flow to the downstream process. The second valve, slaved to the master, should divert excess flow to a bypass loop. This master-slave approach ensures that the pumps and other upstream system components are permitted to operate well within their operating envelope.

Fluid Power Fluid power control valves are used on pneumatic and hydraulic systems or circuits.

Configuration The configuration of fluid power control valves varies with their intended application. The more common configurations include: one-way, two-way, three-way, and four-way.

One-Way One-way valves are typically used for flow and pressure control in fluidpower circuits (see Figure 9.9). Flow-control valves regulate the flow of

Control Valves 191

hydraulic fluid or gases in these systems. Pressure-control valves, in the form of regulators or relief valves, control the amount of pressure transmitted downstream from the valve. In most cases, the types of valves used for flow control are smaller versions of the types of valves used in process control. These include ball, gate, globe, and butterfly valves. Pressure-control valves have a third port to vent excess pressure and prevent it from affecting the downstream piping. The bypass, or exhaust, port has an internal flow-control device, such as a diaphragm or piston, that opens at predetermined set-points to permit the excess pressure to bypass the valve’s primary discharge. In pneumatic circuits, the bypass port vents to the atmosphere. In hydraulic circuits, it must be connected to a piping system that returns to the hydraulic reservoir.

Two-Way A two-way valve has two functional flow-control ports. A two-way, sliding spool directional control valve is shown in Figure 9.10. As the spool moves back and forth, it either allows fluid to flow through the valve or prevents it from flowing. In the open position, the fluid enters the inlet port, flows around the shaft of the spool, and through the outlet port. Because the forces in the cylinder are equal when open, the spool cannot move back and forth. In the closed position, one of the spool’s pistons simply blocks the inlet port, which prevents flow through the valve. A number of features common to most sliding-spool valves are shown in Figure 9.10. The small ports at either end of the valve housing provide a path for fluid that leaks past the spool to flow to the reservoir. This prevents pressure from building up against the ends of the pistons, which would hinder the movement of the spool. When these valves become worn, they may lose balance because of greater leakage on one side of the spool than on

In

In

Open Out

Figure 9.10 Two-way, fluid-power valve

Closed Out

192 Control Valves

#1

#2

A

#3

#1

#2

B

#3

#1

#2

C

#3

Figure 9.11 Three-way, fluid-power valve the other. This can cause the spool to stick as it attempts to move back and forth. Therefore, small grooves are machined around the sliding surface of the piston. In hydraulic valves, leaking liquid encircles the piston, keeping the contacting surfaces lubricated and centered.

Three-Way Three-way valves contain a pressure port, cylinder port, and return or exhaust port (see Figure 9.11). The three-way directional control valve is designed to operate an actuating unit in one direction. It is returned to its original position either by a spring or the load on the actuating unit.

Four-Way Most actuating devices require system pressure in order to operate in two directions. The four-way directional control valve, which contains four ports, is used to control the operation of such devices (see Figure 9.12). The four-way valve also is used in some systems to control the operation of other valves. It is one of the most widely used directional-control valves in fluid-power systems.

Control Valves 193

Air introduced through this passage pushes against the piston which shifts the spool to the right

Centering washers

Springs push against centering washers to center the spool when no air is applied

Pistons seal the air chamber from the hydraulic chamber

Figure 9.12 Four-way, fluid-power valve The typical four-way directional control valve has four ports: pressure port, return port, and two cylinder or work (output) ports. The pressure port is connected to the main system-pressure line, and the return port is connected to the reservoir return line. The two outputs are connected to the actuating unit.

Performance The criteria that determine performance of fluid-power valves are similar to those for process-control valves. As with process-control valves, fluidpower valves must also be selected based on their intended application and function.

Installation When installing fluid power control valves, piping connections are made either directly to the valve body or to a manifold attached to the valve’s base. Care should be taken to ensure that piping is connected to the proper valve port. The schematic diagram that is affixed to the valve body will indicate the proper piping arrangement, as well as the designed operation of

194 Control Valves

the valve. In addition, the ports on most fluid power valves are generally clearly marked to indicate their intended function. In hydraulic circuits, the return or common ports should be connected to a return line that directly connects the valve to the reservoir tank. This return line should not need a pressure-control device, but should have a check valve to prevent reverse flow of the hydraulic fluid. Pneumatic circuits may be vented directly to atmosphere. A return line can be used to reduce noise or any adverse effect that locally vented compressed air might have on the area.

Operating Methods The function and proper operation of a fluid-power valve are relatively simple. Most of these valves have a schematic diagram affixed to the body that clearly explains how to operate the valve.

Valves Figure 9.13 is a schematic of a two-position, cam-operated valve. The primary actuator, or cam, is positioned on the left of the schematic and any secondary actuators are on the right. In this example, the secondary actuator consists of a spring-return and a spring-compensated limit switch. The schematic indicates that when the valve is in the neutral position (right box), flow is directed from the inlet (P) to work port A. When the cam is depressed, the flow momentarily shifts to work port B. The secondary actuator, or spring, automatically returns the valve to its neutral position when the cam returns to its extended position. In these schematics, T indicates the return connection to the reservoir. Figure 9.14 illustrates a typical schematic of a two-position and threeposition directional control valve. The boxes contain flow direction arrows that indicate the flow path in each of the positions. The schematics do not include the actuators used to activate or shift the valves between positions. In a two-position valve, the flow path is always directed to one of the work ports (A or B). In a three-position valve, a third or neutral position is added. In this figure, a Type 2 center position is used. In the neutral position, all ports are blocked, and no flow through the valve is possible. Figure 9.15 is the schematic for the center or neutral position of threeposition directional control valves. Special attention should be given to the type of center position that is used in a hydraulic control valve. When Type 2, 3, and 6 (see Figure 9.15) are used, the upstream side of the valve must

Control Valves 195

“P” “T”

Roller (Cam follower)

A

Push rod trips switch when cam actuates spool

B

Limit switch

Spring holds valve offset in normal operation

A

B

A

B

P

T

P

T

Figure 9.13 Schematic for a cam-operated, two-position valve

have a relief or bypass valve installed. Since the pressure port is blocked, the valve cannot relieve pressure on the upstream side of the valve. The Type 4 center position, called a motor spool, permits the full pressure and volume on the upstream side of the valve to flow directly to the return line and storage reservoir. This is the recommended center position for most hydraulic circuits. The schematic affixed to the valve includes the primary and secondary actuators used to control the valve. Figure 9.16 provides the schematics for three actuator-controlled valves: 1

Double-solenoid, spring-centered, three-position valve

2

Solenoid-operated, spring-return, two-position valve

3

Double-solenoid, detented, two-position valve

196 Control Valves

A

B

P

A

B

T P 2-Position valve B

P

A

T

T

A

B

A

B

P

T

P

T

3-Position valve

Figure 9.14 Schematic of two-position and three-position valves

A

B

A

B

A

B

P T Type 0

P T Type 1

P T Type 2

A

A

A

B

P T Type 3

B

P T Type 4

B

P T Type 6

Figure 9.15 Schematic for center or neutral configurations of three-position valves

Control Valves 197

A

B

A

B

A

B

P

T

P

T

(1)

P

T A

B

A

B

P

T

P

T

(2)

A

B

A

B

P

T

P

T

(3)

Figure 9.16 Actuator-controlled valve schematics

The top schematic, in Figure 9.16, represents a double-solenoid, springcentered, three-position valve. When neither of the two solenoids is energized, the double springs ensure that the valve is in its center or neutral position. In this example, a Type 0 (see Figure 9.15) configuration is used. This neutral-position configuration equalizes the pressure through the valve. Since the pressure port is open to both work ports and the return line, pressure is equalized throughout the system. When the left or primary solenoid is energized, the valve shifts to the left-hand position and directs pressure to work port B. In this position, fluid in the A-side of the circuit returns to the reservoir. As soon as the solenoid is de-energized, the valve shifts back to the neutral or center position. When the secondary (i.e., right) solenoid is energized, the valve redirects flow to port A, and port B returns fluid to the reservoir. The middle schematic, in Figure 9.16, represents a solenoid-operated, spring-return, two-position valve. Unless the solenoid is energized, the pressure port P is connected to work port A. While the solenoid is energized, flow is redirected to work port B. The spring return ensures that the valve is in its neutral (i.e., right) position when the solenoid is de-energized.

198 Control Valves

The bottom schematic, in Figure 9.16, represents a double-solenoid, detented, two-position valve. The solenoids are used to shift the valve between its two positions. A secondary device, called a detent, is used to hold the valve in its last position until the alternate solenoid is energized. Detent configuration varies with the valve type and manufacturer. However, all configurations prevent the valve’s control device from moving until a strong force, such as that provided by the solenoid, overcomes its locking force.

Actuators As with process-control valves, actuators used to control fluid-power valves have a fundamental influence on performance. The actuators must provide positive, real-time response to control inputs. The primary types of actuators used to control fluid-power valves are: mechanical, pilot, and solenoid. Mechanical The use of manually controlled mechanical valves is limited in both pneumatic and hydraulic circuits. Generally, this type of actuator is used only on isolation valves that are activated when the circuit or fluid-power system is shut down for repair or when direct operator input is required to operate one of the system components. Manual control devices (e.g., levers, cams, or palm buttons) can be used as the primary actuator on most fluid power control valves. Normally, these actuators are used in conjunction with a secondary actuator, such as a spring return or detent, to ensure proper operation of the control valve and its circuit. Spring returns are used in applications where the valve is designed to stay open or shut only when the operator holds the manual actuator in a particular position. When the operator releases the manual control, the spring returns the valve to the neutral position. Valves with a detented secondary actuator are designed to remain in the last position selected by the operator until manually moved to another position. A detent actuator is simply a notched device that locks the valve in one of several preselected positions. When the operator applies force to the primary actuator, the valve shifts out of the detent and moves freely until the next detent is reached. Pilot Although there are a variety of pilot actuators used to control fluid-power valves, they all work on the same basic principle. A secondary source of fluid

Control Valves 199

or gas pressure is applied to one side of a sealing device, such as a piston or diaphragm. As long as this secondary pressure remains within preselected limits, the sealing device prevents the control valve’s flow-control mechanism (i.e., spool or poppet) from moving. However, if the pressure falls outside of the preselected window, the actuator shifts and forces the valve’s primary mechanism to move to another position. This type of actuator can be used to sequence the operation of several control valves or operations performed by the fluid-power circuit. For example, a pilot-operated valve is used to sequence the retraction of an airplane’s landing gear. The doors that conceal the landing gear when retracted cannot close until the gear is fully retracted. A pilot-operated valve senses the hydraulic pressure in the gear-retraction circuit. When the hydraulic pressure reaches a pre-selected point that indicates the gear is fully retracted, the pilot-actuated valve triggers the closure circuit for the wheel-well doors. Solenoid Solenoid valves are widely used as actuators for fluid-power systems. This type of actuator consists of a coil that generates an electric field when energized. The magnetic forces generated by this field force a plunger that is attached to the main valve’s control mechanism to move within the coil. This movement changes the position of the main valve. In some applications, the mechanical force generated by the solenoid coil is not sufficient to move the main valve’s control mechanism. When this occurs, the solenoid actuator is used in conjunction with a pilot actuator. The solenoid coil opens the pilot port, which uses system pressure to shift the main valve. Solenoid actuators are always used with a secondary actuator to provide positive control of the main valve. Because of heat buildup, solenoid actuators must be limited to short-duration activation. A brief burst of electrical energy is transmitted to the solenoid’s coil, and the actuation triggers a movement of the main valve’s control mechanism. As soon as the main valve’s position is changed, the energy to the solenoid coil is shut off. This operating characteristic of solenoid actuators is important. For example, a normally closed valve that uses a solenoid actuation can only be open when the solenoid is energized. As soon as the electrical energy is removed from the solenoid’s coil, the valve returns to the closed position. The reverse is true of a normally open valve. The main valve remains open, except when the solenoid is energized.

200 Control Valves

The combination of primary and secondary actuators varies with the specific application. Secondary actuators can be another solenoid or any of the other actuator types that have been previously discussed.

Troubleshooting Although there are limited common control valve failure modes, the dominant problems are usually related to leakage, speed of operation, or complete valve failure. Table 9.1 lists the more common causes of these failures. Table 9.1 Common failure modes of control valves

•

•

Galling

•

• •

Manually actuated

Line pressure too high

•

•

Mechanical damage

•

•

•

•

Not packed properly

•

Packed box too loose

•

Packing too tight

•

•

Threads/lever damaged

•

•

Valve stem bound

•

•

Valve undersized

•

Opens/closes too slow

Excessive wear

Opens/closes too fast

•

Excessive pressure drop

•

Leakage around stem

Leakage through valve

Dirt/debris trapped in valve seat

THE CAUSES

Valve fails to open

Valve fails to close

THE PROBLEM

•

Control Valves 201

Table 9.1 continued

•

Galling

•

•

Mechanical damage (seals, seat)

•

•

•

Pilot port blocked/plugged

•

•

•

Pilot actuated

•

Solenoid actuated

Pilot pressure too high

Opens/closes too slow

•

Opens/closes too fast

Leakage through valve

•

Excessive pressure drop

Valve fails to close


THE CAUSES

Leakage around stem

Valve fails to open

THE PROBLEM

•

Pilot pressure too low

•

•

Corrosion

•

•

•


•

•

•

Galling

•

•

Line pressure too high

•

•

•

Mechanical damage

•

•

•

Solenoid failure

•

•

Solenoid wiring defective

•

•

Wrong type of valve (N-O, N-C)

•

•

•

•

•

Special attention should be given to the valve actuator when conducting a root cause failure analysis. Many of the problems associated with both process and fluid-power control valves are really actuator problems. In particular, remotely controlled valves that use pneumatic, hydraulic, or electrical actuators are subject to actuator failure. In many cases, these

202 Control Valves

failures are the reason a valve fails to properly open, close, or seal. Even with manually controlled valves, the true root cause can be traced to an actuator problem. For example, when a manually operated process-control valve is jammed open or closed, it may cause failure of the valve mechanism. This over-torquing of the valve’s sealing device may cause damage or failure of the seal, or it may freeze the valve stem. Either of these failure modes results in total valve failure.

10 Conveyors Conveyors are used to transport materials from one location to another within a plant or facility. The variety of conveyor systems is almost infinite, but the two major classifications used in typical chemical plants are pneumatic and mechanical. Note that the power requirements of a pneumatic-conveyor system are much greater than for a mechanical conveyor of equal capacity. However, both systems offer some advantages.

Pneumatic Pneumatic conveyors are used to transport dry, free-flowing, granular material in suspension within a pipe or duct. This is accomplished by the use of a high-velocity air stream, or by the energy of expanding compressed air within a comparatively dense column of fluidized or aerated material. Principal uses are: (1) dust collection; (2) conveying soft materials, such as flake or tow; and (3) conveying hard materials, such as fly ash, cement, and sawdust. The primary advantages of pneumatic-conveyor systems are the flexibility of piping configurations and the fact that they greatly reduce the explosion hazard. Pneumatic conveyors can be installed in almost any configuration required to meet the specific application. With the exception of the primary driver, there are no moving parts that can fail or cause injury. However, when used to transport explosive materials, there is still some potential for static charge buildup that could cause an explosion.

Configuration A typical pneumatic-conveyor system consists of Schedule-40 pipe or ductwork, which provides the primary flow path used to transport the conveyed material. Motive power is provided by the primary driver, which can be a fan, fluidizer, or positive-displacement compressor.

Performance Pneumatic conveyor performance is determined by the following factors: (1) primary-driver output; (2) internal surface of the piping or ductwork; and (3) condition of the transported material. Specific factors affecting performance include motive power, friction loss, and flow restrictions.

204 Conveyors

Motive Power The motive power is provided by the primary driver, which generates the gas (typically air) velocity required to transport material within a pneumaticconveyor system. Therefore, the efficiency of the conveying system depends on the primary driver’s operating condition.

Friction Loss Friction loss within a pneumatic-conveyor system is a primary source of efficiency loss. The piping or ductwork must be properly sized to minimize friction without lowering the velocity below the value needed to transport the material.

Flow Restrictions An inherent weakness of pneumatic-conveyor systems is their potential for blockage. The inside surfaces must be clean and free of protrusions or other defects that can restrict or interrupt the flow of material. In addition, when a system is shut down or the velocity drops below the minimum required to keep the transported material suspended, the product will drop out or settle in the piping or ductwork. In most cases, this settled material would compress and lodge in the piping. The restriction caused by this compacted material will reduce flow and eventually result in a complete blockage of the system. Another major contributor to flow restrictions is blockage caused by system backups. This occurs when the end point of the conveyor system (i.e., storage silo, machine, or vessel) cannot accept the entire delivered flow of material. As the transported material backs up in the conveyor piping, it compresses and forms a solid plug that must be manually removed.

Installation All piping and ductwork should be as straight and short as possible. Bends should have a radius of at least three diameters of the pipe or ductwork. The diameter should be selected to minimize friction losses and maintain enough velocity to prevent settling of the conveyed material. Branch lines should be configured to match as closely as possible the primary flow direction and avoid 90-degree angles to the main line. The area of the main conveyor line at any point along its run should be 20 to 25% greater than the sum of all its branch lines. When vertical runs are short in proportion to the horizontal runs, the size of the riser can be restricted to provide

Conveyors 205

additional velocity if needed. If the vertical runs are long, the primary or a secondary driver must provide sufficient velocity to transport the material. Clean-outs, or drop-legs, should be installed at regular intervals throughout the system to permit foreign materials to drop out of the conveyed material. In addition, they provide the means to remove materials that drop out when the system is shut down or air velocity is lost. It is especially important to install adequate clean-out systems near flow restrictions and at the end of the conveyor system.

Operating Methods Pneumatic-conveyor systems must be operated properly to prevent chronic problems, with the primary concern being to maintain constant flow and velocity. If either of these variables is permitted to drop below the system’s design envelope, partial or complete blockage of the conveyor system will occur. Constant velocity can be maintained only when the system is operated within its performance envelope and when regular clean-out is part of the normal operating practice. In addition, the primary driver must be in good operating condition. Any deviation in the primary driver’s efficiency reduces the velocity and can result in partial or complete blockage. The entire pneumatic-conveyor system should be completely evacuated before shutdown to prevent material from settling in the piping or ductwork. In noncontinuous applications, the conveyor system should be operated until all material within the conveyor’s piping is transported to its final destination. Material that is allowed to settle will compact and partially block the piping. Over time, this will cause a total blockage of the conveyor system.

Mechanical There are a variety of mechanical-conveyor systems used in chemical plants. These systems generally are comprised of chain- or screw-type mechanisms.

Chain A commonly used chain-type system is a flight conveyor (e.g., Hefler conveyor), which is used to transport granular, lumpy, or pulverized materials along a horizontal or inclined path within totally enclosed ductwork.

206 Conveyors

The Hefler systems generally have lower power requirements than the pneumatic conveyor and have the added benefit of preventing product contamination. This section focuses primarily on the Hefler-type conveyor because it is one of the most commonly used systems.

Configuration The most common chain conveyor uses a center- or double-chain configuration to provide positive transfer of material within its ductwork. Both chain configurations use hardened bars or U-shaped devices that are an integral part of the chain to drag the conveyed material through the ductwork.

Performance Data used to determine a chain conveyor’s capacity and the size of material that can be conveyed are presented in Table 10.1. Note that these data are for level conveyors. When inclined, capacity data obtained from Table 10.1 must be multiplied by the factors provided in Table 10.2.

Installation The primary installation concerns with Hefler-type conveyor systems are the ductwork and primary-drive system. Ductwork The inside surfaces of the ductwork must be free of defects or protrusions that interfere with the movement of the conveyor’s chain or transported product. This is especially true at the joints. The ductwork must be sized to provide adequate chain clearance but should not be large enough to have areas where the chain-drive bypasses the product. Table 10.1 Approximate capacities of chain conveyors Flight width and depth (inches)

Quantity of material (Ft3 /Ft)

Approximate capacity (short tons/ hour)

Lump size single strand (inches)

Lump size dual strand (inches)

12 × 6 15 × 6 18 × 6 24 × 8 30 × 10 36 × 12

0.40 0.49 0.56 1.16 1.60 2.40

60 73 84 174 240 360

31.5 41.5 5.0

4.0 5.0 6.0 10.0 14.0 16.0

Conveyors 207

Table 10.2 Capacity correction factors for inclined chain conveyors Inclination, degrees Factor

20 0.9

25 0.8

30 0.7

35 0.6

A long horizontal run followed by an upturn is inadvisable because of radial thrust. All bends should have a large radius to permit smooth transition and to prevent material buildup. As with pneumatic conveyors, the ductwork should include clean-out ports at regular intervals for ease of maintenance. Primary-Drive System Most mechanical conveyors use a primary-drive system that consists of an electric motor and a speed-increaser gearbox. The drive-system configuration may vary, depending on the specific application or vendor. However, all configurations should include a single-pointof-failure device, such as a shear pin, to protect the conveyor. The shear pin is critical in this type of conveyor because it is prone to catastrophic failure caused by blockage or obstructions that may lock the chain. Use of the proper shear pin prevents major damage from occurring to the conveyor system. For continuous applications, the primary-drive system must have adequate horsepower to handle a fully loaded conveyor. Horsepower requirements should be determined based on the specific product’s density and the conveyor’s maximum-capacity rating. For intermittent applications, the initial startup torque is substantially greater than for a continuous operation. Therefore, selection of the drive system and the designed failure point of the shear device must be based on the maximum startup torque of a fully loaded system. If either the drive system or designed failure point is not properly sized, this type of conveyor is prone to chronic failures. The predominant failures are frequent breakage of the shear device and trips of the motor’s circuit breaker caused by excessive startup amp loads.

Operating Methods Most mechanical conveyors are designed for continuous operation and may exhibit problems in intermittent-service applications. The primary problem is the startup torque for a fully loaded conveyor. This is especially true

208 Conveyors

for conveyor systems handling material that tends to compact or compress upon settling in a vessel, such as the conveyor trough. The only positive method of preventing excessive startup torque is to ensure that the conveyor is completely empty before shutdown. In most cases, this can be accomplished by isolating the conveyor from its supply for a few minutes prior to shutdown. This time delay permits the conveyor to deliver its entire load of product before it is shut off. In applications where it is impossible to completely evacuate the conveyor prior to shutdown, the only viable option is to jog, or step start, the conveyor. Step starting reduces the amp load on the motor and should control the torque to prevent the shear pin from failing. If instead of step starting, the operator applies full motor load to a stationary, fully loaded conveyor, one of two things will occur: (1) the drive motor’s circuit breaker will trip as a result of excessive amp load, or (2) the shear pin installed to protect the conveyor will fail. Either of these failures adversely impacts production.

Screw The screw, or spiral, conveyor is widely used for pulverized or granular, noncorrosive, nonabrasive materials in systems requiring moderate capacities, distances not more than about 200 feet, and moderate inclines (≤ 35 degrees). It usually costs substantially less than any other type of conveyor and can be made dust-tight by installing a simple cover plate. Abrasive or corrosive materials can be handled with suitable construction of the helix and trough. Conveyors using special materials, such as hardfaced cast iron and linings or coatings, on the components that come into contact with the materials can be specified in these applications. The screw conveyor will handle lumpy material if the lumps are not large in proportion to the diameter of the screw’s helix. Screw conveyors may be inclined. A standard-pitch helix will handle material on inclines up to 35 degrees. Capacity is reduced in inclined applications and Table 10.3 provides the approximate reduction in capacity for various inclines. Table 10.3 Screw conveyor capacity reductions for inclined applications Inclination, degrees Reduction in capacity, %

10 10

15 26

20 45

25 58

30 70

35 78

Conveyors 209

Configuration Screw conveyors have a variety of configurations. Each is designed for specific applications and/or materials. Standard conveyors have a galvanizedsteel rotor, or helix, and trough. For abrasive and corrosive materials (e.g., wet ash), both the helix and trough may be hard-faced cast iron. For abrasives, the outer edge of the helix may be faced with a renewable strip of Stellite (a cobalt alloy produced by Haynes Stellite Co.) or other similarly hard material. Aluminum, bronze, Monel, or stainless steel also may be used to construct the rotor and trough. Short-Pitch Screw The standard helix used for screw conveyors has a pitch approximately equal to its outside diameter. The short-pitch screw is designed for applications with inclines greater than 29 degrees. Variable-Pitch Screw Variable-pitch screws having the short pitch at the feed end automatically control the flow to the conveyor and correctly proportion the load down the screw’s length. Screws having what is referred to as a “short section,” which has either a shorter pitch or smaller diameter, are self-loading and do not require a feeder. Cut-Flight Cut-flight conveyors are used for conveying and mixing cereals, grains, and other light material. They are similar to normal flight or screw conveyors, and the only difference is the configuration of the paddles or screw. Notches are cut in the flights to improve the mixing and conveying efficiency when handling light, dry materials. Ribbon Screw Ribbon screws are used for wet and sticky materials, such as molasses, hot tar, and asphalt. This type of screw prevents the materials from building up and altering the natural frequency of the screw. A buildup can cause resonance problems and possibly catastrophic failure of the unit. Paddle Screw The paddle-screw conveyor is used primarily for mixing materials like mortar and paving mixtures. An example of a typical application is churning ashes and water to eliminate dust.

Performance Process parameters, such as density, viscosity, and temperature, must be constantly maintained within the conveyor’s design operating envelope.

210 Conveyors

Table 10.4 Factor A for self-lubricating bronze bearings Conveyor . diameter, in Factor A

6

9

10

12

14

16

18

20

24

54

96

114

171

255

336

414

510

690

Slight variations can affect performance and reliability. In intermittent applications, extreme care should be taken to fully evacuate the conveyor prior to shutdown. In addition, caution must be exercised when restarting a conveyor in case an improper shutdown was performed and material was allowed to settle. Power Requirements The horsepower requirement for the conveyor-head shaft, H, for horizontal screw conveyors can be determined from the following equation: H = (ALN + CWLF) × 10 − 6 Where: A = Factor for size of conveyor (see Table 10.4) C = Material volume, ft3 /h F = Material factor, unitless (see Table 10.5) L = Length of conveyor, feet N = Conveyor rotation speed (rpm) W = Density of material, lb/ft3 In addition to H, the motor size depends on the drive efficiency (E) and a unitless allowance factor (G), which is a function of H. Values for G are found in Table 10.6. The value for E is usually 90%. Motor hp = HG/E Table 10.5 gives the information needed to estimate the power requirement: percentages of helix loading for five groups of material, maximum material density or capacity, allowable speeds for 6-inch and 20-inch diameter screws, and the factor F.

Conveyors 211

Table 10.5 Power requirements by material group Material group

Max. cross Max. density Max. rpm for section % occupied of material, 6" diameter by the material lb/ft3

Max. rpm for 20" diameter

1 2 3 4 5

45 38 31 25 12 12

110 75 60 50 25

Group 1

F factor is 0.5 for light materials such as barley, beans, brewers grains (dry), coal (pulverized), cornmeal, cottonseed meal, flaxseed, flour, malt, oats, rice, and wheat. Includes fines and granular material. The values of F are: alum (pulverized), 0.6; coal (slack or fines), 0.9; coffee beans, 0.4; sawdust, 0.7; soda ash (light), 0.7; soybeans, 0.5; fly ash, 0.4. Includes materials with small lumps mixed with fines. Values of F are: alum, 1.4; ashes (dry), 4.0; borax, 0.7; brewers grains (wet), 0.6; cottonseed, 0.9; salt, coarse or fine, 1.2; soda ash (heavy), 0.7. Includes semiabrasive materials, fines, granular, and small lumps. Values of F are: acid phosphate (dry), 1.4; bauxite (dry), 1.8; cement (dry), 1.4; clay, 2.0; fuller’s earth, 2.0; lead salts, 1.0; limestone screenings, 2.0; sugar (raw), 1.0; white lead, 1.0; sulfur (lumpy), 0.8; zinc oxide, 1.0. Includes abrasive lumpy materials, which must be kept from contact with hanger bearings. Values of F are: wet ashes, 5.0; flue dirt, 4.0; quartz (pulverized), 2.5; silica sand, 2.0; sewage sludge (wet and sandy), 6.0.

Group 2

Group 3

Group 4

Group 5

50 50 75 100

170 120 90 70 30

Table 10.6 Allowance factor H, hp G

1 2

1–2 1.5

2–4 1.25

4–5 1.1

5 1.0

Volumetric Efficiency Screw-conveyor performance is also determined by the volumetric efficiency of the system. This efficiency is determined by the amount of slip or bypass generated by the conveyor. The amount of slip in a screw

212 Conveyors

conveyor is primarily determined by three factors: product properties, screw efficiency, and clearance between the screw and the conveyor barrel or housing. Product Properties Not all materials or products have the same flow characteristics. Some have plastic characteristics and flow easily. Others do not self-adhere and tend to separate when pumped or mechanically conveyed. As a result, the volumetric efficiency is directly affected by the properties of each product. This also impacts screw performance. Screw Efficiency Each of the common screw configurations (i.e., short pitch, variable pitch, cut flights, ribbon, and paddle) has varying volumetric efficiencies, depending on the type of product that is conveyed. Screw designs or configurations must be carefully matched to the product to be handled by the system. For most medium- to high-density products in a chemical plant, the variablepitch design normally provides the highest volumetric efficiency and lowest required horsepower. Cut-flight conveyors are highly efficient for light, nonadhering products, such as cereals, but are inefficient when handling heavy, cohesive products. Ribbon conveyors are used to convey heavy liquids, such as molasses, but are not very efficient and have a high slip ratio. Clearance Improper clearance is the source of many volumetric-efficiency problems. It is important to maintain proper clearance between the outer ring, or diameter, of the screw and the conveyor’s barrel, or housing, throughout the operating life of the conveyor. Periodic adjustments to compensate for wear, variations in product, and changes in temperature are essential. While the recommended clearance varies with specific conveyor design and the product to be conveyed, excessive clearance severely impacts conveyor performance as well.

Installation Installation requirements vary greatly with screw-conveyor design. The vendor’s Operating and Maintenance (O&M) manuals should be consulted and followed to ensure proper installation. However, as with practically all mechanical equipment, there are basic installation requirements common

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to all screw conveyors. Installation requirements presented here should be evaluated in conjunction with the vendor’s O&M manual. If the information provided here conflicts with the vendor-supplied information, the O&M manual’s recommendations should always be followed. Foundation The conveyor and its support structure must be installed on a rigid foundation that absorbs the torsional energy generated by the rotating screws. Because of the total overall length of most screw conveyors, a single foundation that supports the entire length and width should be used. There must be enough lateral (i.e., width) stiffness to prevent flexing during normal operation. Mounting conveyor systems on decking or suspended-concrete flooring should provide adequate support. Support Structure Most screw conveyors are mounted above the foundation level on a support structure that generally has a slight downward slope from the feed end to the discharge end. While this improves the operating efficiency of the conveyor, it also may cause premature wear of the conveyor and its components. The support’s structural members (i.e., I-beams and channels) must be adequately rigid to prevent conveyor flexing or distortion during normal operation. Design, sizing, and installation of the support structure must guarantee rigid support over the full operating range of the conveyor. When evaluating the structural requirements, variations in product type, density, and operating temperature must also be considered. Since these variables directly affect the torsional energy generated by the conveyor, the worst-case scenario should be used to design the conveyor’s support structure. Product-Feed System One of the major limiting factors of screw conveyors is their ability to provide a continuous supply of incoming product. While some conveyor designs, such as those having a variable-pitch screw, provide the ability to self-feed, their installation should include a means of ensuring a constant, consistent incoming supply of product. In addition, the product-feed system must prevent entrainment of contaminants in the incoming product. Normally, this requires an enclosure that seals the product from outside contaminants.

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Operating Methods As previously discussed, screw conveyors are sensitive to variations in incoming product properties and the operating environment. Therefore, the primary operating concern is to maintain a uniform operating envelope at all times, in particular by controlling variations in incoming product and operating environment. Incoming-Product Variations Any measurable change in the properties of the incoming product directly affects the performance of a screw conveyor. Therefore, the operating practices should limit variations in product density, temperature, and viscosity. If they occur, the conveyor’s speed should be adjusted to compensate for them. For property changes directly related to product temperature, preheaters or coolers can be used in the incoming-feed hopper, and heating/cooling traces can be used on the conveyor’s barrel. These systems provide a means of achieving optimum conveyor performance despite variations in incoming product. Operating-Environment Variations Changes in the ambient conditions surrounding the conveyor system may also cause deviations in performance. A controlled environment will substantially improve the conveyor’s efficiency and overall performance. Therefore, operating practices should include ways to adjust conveyor speed and output to compensate for variations. The conveyor should be protected from wind chill, radical variations in temperature and humidity, and any other environment-related variables.

11 Couplings Couplings are designed to provide two functions: (1) to transmit torsional power between a power source and driven unit and (2) to absorb torsional variations in the drive train. They are not designed to correct misalignment between two shafts. While certain types of couplings provide some correction for slight misalignment, reliance on these devices to obtain alignment is not recommended.

Coupling Types The sections to follow provide overviews of the more common coupling types: rigid and flexible. Also discussed are couplings used for special applications: floating-shaft (spacer) and fluid (hydraulic).

Rigid Couplings A rigid coupling permits neither axial nor radial relative motion between the shafts of the driver and driven unit. When the two shafts are connected solidly and properly, they operate as a single shaft. A rigid coupling is primarily used for vertical applications, e.g., vertical pumps. Types of rigid couplings discussed in this section are flanged, split, and compression. Flanged couplings are used where there is free access to both shafts. Split couplings are used where access is limited on one side. Both flanged and split couplings require the use of keys and keyways. Compression couplings are used when it is not possible to use keys and keyways.

Flanged Couplings A flanged rigid coupling is comprised of two halves, one located on the end of the driver shaft and the other on the end of the driven shaft. These halves are bolted together to form a solid connection. To positively transmit torque, the coupling incorporates axially fitted keys and split circular key rings or dowels, which eliminate frictional dependency for transmission. The use of flanged couplings is restricted primarily to vertical pump shafts. A typical flanged rigid coupling is illustrated in Figure 11.1.

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Figure 11.1 Typical flanged rigid coupling

Split Couplings A split rigid coupling, also referred to as a clamp coupling, is basically a sleeve that is split horizontally along the shaft and held together with bolts. It is clamped over the adjoining ends of the driver and driven shafts, forming a solid connection. Clamp couplings are used primarily on vertical pump shafting. A typical split rigid coupling is illustrated in Figure 11.2. As with the flanged coupling, the split rigid coupling incorporates axially fitted keys and split circular key rings to eliminate frictional dependency in the transmission of torque.

Compression Coupling A rigid compression coupling is comprised of three pieces: a compressible core and two encompassing coupling halves that apply force to the core. The core is comprised of a slotted bushing that has been machine bored to fit both ends of the shafts. It also has been machined with a taper on its external diameter from the center outward to both ends. The coupling halves are finish bored to fit this taper. When the coupling halves are bolted together, the core is compressed down on the shaft by the two halves, and the resulting frictional grip transmits the torque without the use of keys. A typical compression coupling is illustrated in Figure 11.3.

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Figure 11.2 Typical split rigid coupling

Figure 11.3 Typical compression rigid coupling

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Flexible Couplings Flexible couplings, which are classified as mechanical flexing, material flexing, or combination, allow the coupled shafts to slide or move relative to each other. Although clearances are provided to permit movement within specified tolerance limits, flexible couplings are not designed to compensate for major misalignments. (Shafts must be aligned to less than 0.002 inches for proper operation.) Significant misalignment creates a whipping movement of the shaft, adds thrust to the shaft and bearings, causes axial vibrations, and leads to premature wear or failure of equipment.

Mechanical Flexing Mechanical-flexing couplings provide a flexible connection by permitting the coupling components to move or slide relative to each other. In order to permit such movement, clearance must be provided within specified limits. It is important to keep cross loading on the connected shafts at a minimum. This is accomplished by providing adequate lubrication to reduce wear on the coupling components. The most popular of the mechanical-flexing type are the chain and gear couplings.

Chain Chain couplings provide a good means of transmitting proportionately high torque at low speeds. Minor shaft misalignment is compensated for by means of clearances between the chain and sprocket teeth and the clearance that exists within the chain itself. The design consists of two hubs with sprocket teeth connected by a chain of the single-roller, double-roller, or silent type. A typical example of a chain coupling is illustrated in Figure 11.4. Special-purpose components may be specified when enhanced flexibility and reduced wear is required. Hardened sprocket teeth, special tooth design, and barrel-shaped rollers are available for special needs. Lightduty drives are sometimes supplied with nonmetallic chains on which no lubrication should be used.

Gear Gear couplings are capable of transmitting proportionately high torque at both high and low speeds. The most common type of gear coupling consists of two identical hubs with external gear teeth and a sleeve, or cover, with matching internal gear teeth. Torque is transmitted through the gear teeth,

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Roller-chain coupling.

Coupling cover (1/2 shown) (optional) Roller chain 1 req’d. to join couplers

Coupling body(s) 1 req’d. for each shaft

Figure 11.4 Typical chain coupling whereas the necessary sliding action and ability for slight adjustments in position comes from a certain freedom of action provided between the two sets of teeth. Slight shaft misalignment is compensated for by the clearance between the matching gear teeth. However, any degree of misalignment decreases the useful life of the coupling and may cause damage to other machine-train components such as bearings. A typical example of a gear-tooth coupling is illustrated in Figure 11.5.

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Figure 11.5 Typical gear-tooth coupling

Material-Flexing Material-flexing couplings incorporate elements that accommodate a certain amount of bending or flexing. The material-flexing group includes laminated disk-ring, bellows, flexible shaft, diaphragm, and elastomeric couplings. Various materials, such as metal, plastics, or rubber, are used to make the flexing elements in these couplings. The use of the couplings is governed by the operational fatigue limits of these materials. Practically all metals have fatigue limits that are predictable; therefore, they permit definite boundaries of operation to be established. Elastomers such as plastic or rubber, however, usually do not have a well defined fatigue limit. Their service life is determined primarily by conditions of installation and operation.

Laminated Disk-Ring The laminated disk-ring coupling consists of shaft hubs connected to a single flexible disk, or a series of disks, that allows axial movement. The laminated disk-ring coupling also reduces heat and axial vibration that can transmit

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Morflex couplings

Laminated disk-ring coupling (standard double-engagement)

Dropout style

Laminated disk-ring coupling (high speed spacer type)

Figure 11.6 Typical laminated disk-ring couplings between the driver and driven unit. Figure 11.6 illustrates some typical laminated disk-ring couplings.

Bellows Bellows couplings consist of two shaft hubs connected to a flexible bellows. This design, which compensates for minor misalignment, is used at moderate rotational torque and shaft speed. This type of coupling provides flexibility to compensate for axial movement and misalignment caused by thermal expansion of the equipment components. Figure 11.7 illustrates a typical bellows coupling.

Flexible Shaft or Spring Flexible shaft or spring couplings are generally used in small equipment applications that do not experience high torque loads. Figure 11.8 illustrates a typical flexible shaft coupling.

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Figure 11.7 Typical bellows coupling

Figure 11.8 Typical flexible shaft coupling

Diaphragm Diaphragm couplings provide torsional stiffness while allowing flexibility in axial movement. Typical construction consists of shaft hub flanges and a diaphragm spool, which provides the connection between the driver and driven unit. The diaphragm spool normally consists of a center shaft fastened to the inner diameter of a diaphragm on each end of the spool shaft. The shaft hub flanges are fastened to the outer diameter of the diaphragms to complete the mechanical connection. A typical diaphragm coupling is illustrated in Figure 11.9.

Elastomeric Elastomeric couplings consist of two hubs connected by an elastomeric element. The couplings fall into two basic categories, one with the element

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Figure 11.9 Typical diaphragm coupling

placed in shear and the other with the element placed in compression. The coupling compensates for minor misalignments because of the flexing capability of the elastomer. These couplings are usually applied in light- or medium-duty applications running at moderate speeds. With the shear-type coupling, the elastomeric element may be clamped or bonded in place, or fitted securely to the hubs. The compression-type couplings may be fitted with projecting pins, bolts, or lugs to connect the components. Polyurethane, rubber, neoprene, or cloth and fiber materials are used in the manufacture of these elements. Although elastomeric couplings are practically maintenance free, it is good practice to periodically inspect the condition of the elastomer and the alignment of the equipment. If the element shows signs of defects or wear, it should be replaced and the equipment realigned to the manufacturer’s specifications. Typical elastomeric couplings are illustrated in Figure 11.10.

Combination (Metallic-Grid) The metallic-grid coupling is an example of a combination of mechanicalflexing and material-flexing type couplings. Typical metallic-grid couplings are illustrated in Figure 11.11.

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Figure 11.10 Typical elastomeric couplings The metallic-grid coupling is a compact unit capable of transmitting high torque at moderate speeds. The construction of the coupling consists of two flanged hubs, each with specially grooved slots cut axially on the outer edges of the hub flanges. The flanges are connected by means of a serpentineshaped spring grid that fits into the grooved slots. The flexibility of this grid provides torsional resilience.

Special Application Couplings Two special application couplings are discussed in this section: (1) floatingshaft or spacer coupling and (2) hydraulic or fluid coupling.

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Figure 11.11 Typical metallic-grid couplings

Floating-Shaft or Spacer Coupling Regular flexible couplings connect the driver and driven shafts with relatively close ends and are suitable for limited misalignment. However, allowances sometimes have to be made to accommodate greater misalignment or when the ends of the driver and driven shafts have to be separated by a considerable distance. Such is the case, for example, with end-suction pump designs in which the power unit of the pump assembly is removed for maintenance by being axially moved toward the driver. If neither the pump nor the driver can be readily removed, they should be separated sufficiently to permit withdrawal of the pump’s power unit. An easily removable flexible coupling of sufficient length (i.e., floating-shaft or spacer coupling) is required for this type of maintenance. Examples of couplings for this type of application are shown in Figure 11.12. In addition to the maintenance application described above, this coupling (also referred to as extension or spacer sleeve coupling) is commonly used where equipment is subject to thermal expansion and possible misalignment because of high process temperatures. The purpose of this type of coupling is to prevent harmful misalignment with minimum separation of the driver and driven shaft ends. An example of a typical floating-shaft coupling for this application is shown in Figure 11.13. The floating-shaft coupling consists of two support elements connected by a shaft. Manufacturers use various approaches in their designs for these couplings. For example, each of the two support elements may be of the

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Laminated disk-ring coupling, spacer type

Gear coupling, spindle type

Gear coupling, high speed spacer type

Figure 11.12 Typical floating-shaft or spacer couplings

Figure 11.13 Typical floating-shaft or spacer couplings for high-temperature applications

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Figure 11.14 Typical hydraulic coupling

single-engagement type, may consist of a flexible half-coupling on one end and a rigid half-coupling on the other end, or may be completely flexible with some piloting or guiding supports. Floating-shaft gear couplings usually consist of a standard coupling with a two-piece sleeve. The sleeve halves are bolted to rigid flanges to form two single-flex couplings. An intermediate shaft that permits the transmission of power between widely separated drive components, in turn, connects these.

Hydraulic or Fluid Hydraulic couplings provide a soft start with gradual acceleration and limited maximum torque for fixed operating speeds. Hydraulic couplings are typically used in applications that undergo torsional shock from sudden changes in equipment loads (e.g., compressors). Figure 11.14 is an illustration of a typical hydraulic coupling.

Coupling Selection Periodically, worn or broken couplings must be replaced. One of the most important steps in performing this maintenance procedure is to ensure that the correct replacement parts are used. After having determined the cause of failure, it is crucial to identify the correct type and size of coupling needed.

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Even if practically identical in appearance to the original, a part still may not be an adequate replacement. The manufacturer’s specification number usually provides the information needed for part selection. If the part is not in stock, a cross-reference guide will provide the information needed to verify ratings and to identify a coupling that meets the same requirements as the original. Criteria that must be considered in part selection include: equipment type, mode of operation, and cost. Each of these criteria is discussed in the sections to follow.

Equipment Type Coupling selection should be application specific and, therefore, it is important to consider the type of equipment that it connects. For example, demanding applications such as variable, high-torque machine trains require couplings that are specifically designed to absorb radical changes in speed and torque (e.g., metallic-grid). Less demanding applications such as runout table rolls can generally get by with elastomeric couplings. Table 11.1 lists the coupling type commonly used in a particular application.

Mode of Operation Coupling selection is highly dependent on the mode of operation, which includes torsional characteristics, speed, and the operating envelope.

Torsional Characteristics Torque requirements are a primary concern during the selection process. In all applications in which variable or high torque is transmitted from the driver to the driven unit, a flexible coupling rated for the maximum torque requirement must be used. Rigid couplings are not designed to absorb variations in torque and should not be used.

Speed Two speed-related factors should be considered as part of the selection process: maximum speed and speed variation. Maximum Speed When selecting coupling type and size, the maximum speed rating must be considered, which can be determined from the vendor’s catalog.

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Table 11.1 Coupling application overview Application

Coupling selection recommendation

Limited Misalignment Compensation Variable, high-torque machine trains operating at moderate speeds Run-out table rolls Vertical pump shafting Keys and keyways not appropriate (e.g., brass shafts) Transmission of proportionately high torque at low speeds Transmission of proportionately high torque at both high and low speeds Allowance for axial movement and reduction of heat and axial vibration Moderate rotational torque and shaft speed Small equipment that does not experience high torque loads Torsional stiffness while allowing flexibility in axial movement Light- or medium-duty applications running at moderate speeds Gradual acceleration and limited maximum torque for fixed operating speeds (e.g., compressors) Variable or high torque and/or speed transmission

Metallic-grid combination couplings Elastomeric flexible couplings Flanged rigid couplings, split rigid or clamp couplings Rigid compression couplings Chain couplings (mechanical-flexing) Gear couplings (mechanical-flexing) Laminated disk-ring couplings (material-flexing) Bellows couplings (material-flexing) Flexible shaft or spring couplings (material-flexing) Diaphragm material-flexing couplings Elastomeric couplings (materialflexing) Hydraulic or fluid couplings

Flexible couplings rated for the maximum torque requirement

Greater Misalignment Compensation Maintenance requiring considerable distance between the driver and driven shaft ends. Misalignment results from expansion due to high process temperatures.

Floating-shaft or spacer couplings

Note: Rigid couplings are not designed to absorb variations in torque and speed and should not be used in such applications. Maximum in-service coupling speed should be at least 15% below the maximum coupling speed rating.

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The maximum in-service speed of a coupling should be well below (at least 15%) the maximum speed rating. The 15% margin provides a service factor that should be sufficient to prevent coupling damage or catastrophic failure. Speed Variation Variation in speed equates to a corresponding variation in torque. Most variable-speed applications require some type of flexible coupling capable of absorbing these torsional variations.

Operating Envelope The operating envelope defines the physical requirements, dimensions, and type of coupling needed in a specific application. The envelope information should include: shaft sizes, orientation of shafts, required horsepower, full range of operating torque, speed ramp rates, and any other data that would directly or indirectly affect the coupling.

Cost Coupling cost should not be the deciding factor in the selection process, although it will certainly play a part in it. Although higher performance couplings may be more expensive, they actually may be the cost-effective solution in a particular application. Selecting the most appropriate coupling for an application not only extends coupling life, but also improves the overall performance of the machine train and its reliability.

Installation Couplings must be installed properly if they are to operate satisfactorily. This section discusses shaft and coupling preparation, coupling installation, and alignment.

Shaft Preparation A careful inspection of both shaft ends must be made to ensure that no burrs, nicks, or scratches are present that will damage the hubs. Potentially damaging conditions must be corrected before coupling installation. Emery cloth should be used to remove any burrs, scratches, or oxidation that may be present. A light film of oil should be applied to the shafts prior to installation.

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Keys and keyways also should be checked for similar defects and to ensure that the keys fit properly. Properly sized key stock must be used with all keyways; do not use bar stock or other material.

Coupling Preparation The coupling must be disassembled and inspected prior to installation. The location and position of each component should be noted so that it can be reinstalled in the correct order. When old couplings are removed for inspection, bolts and bolt holes should be numbered so that they can be installed in the same location when the coupling is returned to service. Any defects, such as burrs, should be corrected before the coupling is installed. Defects on the mating parts of the coupling can cause interference between the bore and shaft, preventing proper operation of the coupling.

Coupling Installation Once the inspection shows the coupling parts to be free of defects, the hubs can be mounted on their respective shafts. If it is necessary to heat the hubs to achieve the proper interference fit, an oil or water bath should be used. Spot heating using a flame or torch should be avoided, as it causes distortion and may adversely affect the hubs. Care must be exercised during installation of a new coupling or the reassembly of an existing unit. Keys and keyways should be coated with a sealing compound that is resistant to the lubricant used in the coupling. Seals should be inspected to ensure that they are pliable and in good condition. They must be installed properly in the sleeve with the lip in good contact with the hub. Sleeve flange gaskets must be whole, in good condition, clean, and free of nicks or cracks. Lubrication plugs must be cleaned before being installed and must fit tightly. The specific installation procedure is dependent on the type and mounting configuration of the coupling. However, common elements of all coupling installations include: spacing, bolting, lubrication, and the use of matching parts. The sections that follow discuss these installation elements.

Spacing Spacing between the mating parts of the coupling must be within manufacturer’s tolerances. For example, an elastomeric coupling must have a specific distance between the coupling faces. This distance determines the

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position of the rubber boot that provides transmission of power from the driver to the driven machine component. If this distance is not exact, the elastomer will attempt to return to its relaxed position, inducing excessive axial movement in both shafts.

Bolting Couplings are designed to use a specific type of bolt. Coupling bolts have a hardened cylindrical body sized to match the assembled coupling width. Hardened bolts are required because standard bolts do not have the tensile strength to absorb the torsional and shearing loads in coupling applications and may fail, resulting in coupling failure and machine-train damage.

Lubrication Most couplings require lubrication, and care must be taken to ensure that the proper type and quantity is used during the installation process. Inadequate or improper lubrication reduces coupling reliability and reduces its useful life. In addition, improper lubrication can cause serious damage to the machine train. For example, when a gear-type coupling is overfilled with grease, the coupling will lock. In most cases, its locked position will increase the vibration level and induce an abnormal loading on the bearings of both the driver and driven unit, resulting in bearing failure.

Matching Parts Couplings are designed for a specific range of applications, and proper performance depends on the total design of the coupling system. As a result, it is generally not a good practice to mix coupling types. Note, however, that it is common practice in some steel industry applications to use coupling halves from two different types of couplings. For example, a rigid coupling half is sometimes mated to a flexible coupling half, creating a hybrid. While this approach may provide short-term power transmission, it can result in an increase in the number, frequency, and severity of machine-train problems.

Coupling Alignment The last step in the installation process is verifying coupling and shaft alignment. With the exception of special application couplings such as spindles and jackshafts, all couplings must be aligned within relatively close tolerances (i.e., 0.001 to 0.002 inch).

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Lubrication and Maintenance Couplings require regular lubrication and maintenance to ensure optimum trouble-free service life. When proper maintenance is not conducted, premature coupling failure and/or damage to machine-train components such as bearings can be expected.

Determining Cause of Failure When a coupling failure occurs, it is important to determine the cause of failure. Failure may result from a coupling defect, an external condition, or workmanship during installation. Most faults are attributed to poorly machined surfaces causing out-ofspecification tolerances, although defective material failures also occur. Inadequate material hardness and poor strength factors contribute to many premature failures. Other common causes are improper coupling selection, improper installation, and/or excessive misalignment.

Lubrication Requirements Lubrication requirements vary depending on application and coupling type. Because rigid couplings do not require lubrication, this section discusses lubrication requirements for mechanical-flexing, material-flexing, and combination flexible couplings only.

Mechanical-Flexing Couplings It is important to follow the manufacturer’s instructions for lubricating mechanical-flexing couplings, which must be lubricated internally. Lubricant seals must be in good condition and properly fitted into place. Coupling covers contain the lubricant and prevent contaminants from entering the coupling interior. The covers are designed in two configurations, split either horizontally or vertically. Holes are provided in the covers to allow lubricant to be added without coupling disassembly. Gear couplings are one type of mechanical-flexing coupling, and there are several ways to lubricate them: grease pack, oil fill, oil collect, and continuous oil flow. Either grease or oil can be used at speeds of 3,600 rpm to 6,000 rpm. Oil is normally used as the lubricant in couplings operating over 6,000 rpm. Grease and oil-lubricated units have end gaskets and seals, which

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are used to contain the lubricant and seal out the entry of contaminants. The sleeves have lubrication holes, which permit flushing and relubrication without disturbing the sleeve gasket or seals.

Material-Flexing Couplings Material-flexing couplings are designed to be lubrication free.

Combination Couplings Combination (metallic-grid) couplings are lubricated in the same manner as mechanical-flexing couplings.

Periodic Inspections It is important to perform periodic inspections of all mechanical equipment and systems that incorporate rotating parts, including couplings and clutches.

Mechanical-Flexing Couplings To maintain coupling reliability, mechanical-flexing couplings require periodic inspections on a time- or condition-based frequency established by the history of the equipment’s coupling life or a schedule established by the predictive maintenance engineer. Items to be included in an inspection are listed below. If any of these items or conditions is discovered, the coupling should be evaluated to determine its remaining operational life or repaired/replaced. ●

Inspect lubricant for traces of metal (indicating component wear).

●

Visually inspect coupling mechanical components (roller chains and gear teeth, and grid members) for wear and/or fatigue.

●

Inspect seals to ensure they are pliable and in good condition. They must be installed properly in the sleeve with the lip in good contact with the hub.

●

Sleeve flange gaskets must be whole, in good condition, clean, and free of nicks or cracks.

●

Lubrication plugs must be clean (to prevent the introduction of contaminants to the lubricant and machine surfaces) before being installed and must be torqued to the manufacturer’s specifications.

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●

Setscrews and retainers must be in place and tightened to manufacturer’s specifications.

●

Inspect shaft hubs, keyways, and keys for cracks, breaks, and physical damage.

●

Under operating conditions, perform thermographic scans to determine temperature differences on the coupling (indicates misalignment and/or uneven mechanical forces).

Material-Flexing Couplings Although designed to be lubrication-free, material-flexing couplings also require periodic inspection and maintenance. This is necessary to ensure that the coupling components are within acceptable specification limits. Periodic inspections for the following conditions are required to maintain coupling reliability. If any of these conditions are found, the coupling should be evaluated to determine its remaining operational life or repaired/replaced. ●

Inspect flexing element for signs of wear or fatigue (cracks, element dust, or particles).

●

Setscrews and retainers must be in place and tightened to manufacturer’s specifications.

●

Inspect shaft hubs, keyways, and keys for cracks, breaks, and physical damage.

●

Under operating conditions, perform thermographic scans for temperature differences on the coupling, which indicates misalignment and/or uneven mechanical forces.

Combination Couplings Mechanical components (e.g., grid members) should be visually inspected for wear and/or fatigue. In addition to the items for mechanical-flexing couplings, the grid members on metallic-grid couplings should be replaced if any signs of wear are observed.

Rigid Couplings The mechanical components of rigid couplings (e.g., hubs, bolts, compression sleeves and halves, keyways, and keys) should be visually inspected

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for cracks, breaks, physical damage, wear, and/or fatigue. Any component having any of these conditions should be replaced.

Keys, Keyways, and Keyseats A key is a piece of material, usually metal, placed in machined slots or grooves cut into two axially oriented parts in order to mechanically lock them together. For example, keys are used in making the coupling connection between the shaft of a driver and a hub or flange on that shaft. Any rotating element whose shaft incorporates such a keyed connection is referred to as a keyed-shaft rotor. Keys provide a positive means for transmitting torque between the shaft and coupling hub when a key is properly fitted in the axial groove. The groove into which a key is fitted is referred to as a keyseat when referring to shafts, and a keyway when referring to hubs. Keyseating is the actual machine operation of producing keyseats. Keyways are normally made on a keys eater or by a broach. Keyseats are normally made with a rotary or end mill cutter. Figure 11.15 is an example of a keyed shaft that shows the key size versus the shaft diameter. Because of standardization and interchangeability, keys are generally proportioned with relation to shaft diameter instead of torsional load. The effective key length, “L” is that portion of key having full bearing on hub and shaft. Note that the curved portion of the keyseat made with a

Top view

Top view

L

L

Rotary cutter

End mill cutter

Figure 11.15 Keyed shaft: key size versus shaft diameter

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TOP VIEWS

Square ends

SIDE VIEWS

Plain taper

Square and round

Rounded ends

Gib head taper

Figure 11.16 Key shapes rotary cutter does not provide full key bearing, so “L” does not include this distance. The use of an end mill cutter results in a square-ended keyseat. Figure 11.16 shows various key shapes: square ends, one square end and one round end, rounded ends, plain taper, and gib head taper. The majority of keys are square in cross section, which are preferred through 4 12 " diameter shafts. For bores over 4 12 " and thin wall sections of hubs, the rectangular (flat) key is used. The ends are either square, rounded or gib-head. The gib-head is usually used with taper keys. If special considerations dictate the use of a keyway in the hub shallower than the preferred square key, it is recommended that the standard rectangular (flat) key be used. Hub bores are usually straight, although for some special applications taper bores are sometimes specified. For smaller diameters, bores are designed for clearance fits, and a setscrew is used over the key. The major advantage of a clearance fit is that hubs can be easily assembled and disassembled. For larger diameters, the bores are designed for interference fits without setscrews. For rapid-reversing applications, interference fits are required. The sections to follow discuss determining keyway depth and width, keyway manufacturing tolerances, key stress calculations, and shaft stress calculations.

Determining Keyway Depth and Width The formula given below, along with Figure 11.17, Table 11.2 (square keys), and Table 11.3 (flat keys), illustrates how the depth and width of standard

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Shafts

Hubs H /2

Y H/2

Y

W

D S

T

D

S = D⫺Y⫺H 2

T = D⫺Y⫹H⫹C 2

Figure 11.17 Shaft and hub dimensions

square and flat keys and keyways for shafts and hubs are determined. D − D2 − W 2 Y = 2 Where: C = Allowance or clearance for key, inches D = Nominal shaft or bore diameter, inches H = Nominal key height, inches W = Nominal key width, inches Y = Chordal height, inches Note: Tables 11.2 and 11.3 shown below are prepared for manufacturing use. Dimensions given are for standard shafts and keyways.

Keyway Manufacturing Tolerances Keyway manufacturing tolerances (illustrated in Figure 11.18) are referred to as offset (centrality) and lead (cross axis). Offset or centrality is referred

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Table 11.2 Standard square keys and keyways (inches)* Keyways Diameter of holes (inclusive)

Width

Depth

Key stock

5/16 to 7/16 1/2 to 9/16 5/8 to 7/8 1 5/16 to 11/4 1 5/16 to 13/8 1 7/16 to 13/4 1 13/16 to 21/4 2 5/16 to 23/4 2 13/16 to 31/4 3 5/16 to 33/4 3 13/16 to 4 1/2

3/32 1/8 3/16 1/4 5/16 3/8 1/2 5/8 3/4 7/8 1

3/64 1/16 3/32 1/8 5/32 3/16 1/4 5/16 3/8 7/16 1/2

3/32 × 3/32 1/8 × 1/8 3/16 × 3/16 1/4 × 1/4 5/16 × 5/16 3/8 × 3/8 1/2 × 1/2 5/8 × 5/8 3/4 × 3/4 7/8 × 7/8 1×1

Source: The Falk Corporation *Square keys are normally used through shaft diameter 4 12 "; larger shafts normally use flat keys.

to as dimension “N”; lead or cross axis is referred to as dimension “J.” Both must be kept within permissible tolerances, usually 0.002 inches.

Key Stress Calculations Calculations for shear and compressive key stresses are based on the following assumptions: 1

The force acts at the radius of the shaft.

2

The force is uniformly distributed along the key length.

3

None of the tangential load is carried by the frictional fit between shaft and bore.

The shear and compressive stresses in a key are calculated using the following equations (see Figure 11.19): Ss =

2T (d)x(w)x(L)

Sc =

2T (d)x(h1 )x(L)

240 Couplings

Table 11.3 Standard flat keys and keyways (inches) Keyways Diameter of holes (inclusive)

Width

Depth

Key stock

1/2 to 9/16" 5/8 to 7/8" 1 5/16 to 1 1/4" 1 5/16 to 1 3/8" 1 7/16 to 1 3/4" 1 13/16 to 2 1/4" 2 5/16 to 2 3/4" 2 13/16 to 3 1/4" 3 5/16 to 3 3/4" 3 13/16 to 4 1/2" 4 9/16 to 5 1/2" 5 9/16 to 6 1/2" 6 9/16 to 7 1/2" 7 9/16 to 9" 9 1/16 to 11" 11 1/16 to 13" 13 1/16 to 15" 15 1/16 to 18" 18 1/16 to 22" 22 1/16 to 26"

1/8 3/16 1/4 5/16 3/8 1/2 5/8 3/4 7/8 1 1 1/4 1 1/2 1 3/4 2 2 1/2 3 3 1/2 4 5 6

3/64 1/16 3/32 1/8 1/8 3/16 7/32 1/4 5/16 3/8 7/16 1/2 5/8 3/4 7/8 1 1 1/4 1 1/2 1 3/4 4

1/8 × 1/32 3/16 × 1/8 1/4 × 3/16 5/16 × 1/4 3/8 × 1/4 1/2 × 3/8 5/8 × 7/16 3/4 × 1/2 7/8 × 5/8 1 × 3/4 1 1/4 × 7/8 1 1/2 × 1 1 3/4 × 1 1/4 2 × 1 1/2 2 1/2 × 1 3/4 3×2 3 1/2 × 2 1/2 4×3 5 × 3 1/2

26 1/16 to 30"

7

5

Source: The Falk Corporation

Where: d = Shaft diameter, inches (use average diameter for taper shafts) h1 = Height of key in the shaft or hub that bears against the keyway, inches. Should equal h2 for square keys. For designs where unequal portions of the key are in the hub or shaft, h1 is the minimum portion.

Couplings 241

(a)

Offset or centrality C C Bore

Shaft N

C

C Keyseat

(b)

Lead or cross axis C Keyseat C Shaft C Bore

C Keyseat

Lead Lead

Figure 11.18 Manufacturing tolerances: offset and lead hp = Power, horsepower L = Effective length of key, inches rpm = Revolutions per minute Ss = Shear stress, psi Sc = Compressive stress, psi T = Shaft torque, lb-in or w = Key width, inches

hp × 63000 rpm

242 Couplings

h1 H h2

W

Clearance

d Shaft diameter

Figure 11.19 Measurements used in calculating shear and compressive key stress Table 11.4 Allowable stresses for AISI 1018 and AISI 1045 Allowable stresses - psi

Material

Heat treatment

Shear

Compressive

AISI 1018 AISI 1045

None 255–300 Bhn

7,500 15,000

15,000 30,000

Source: The Falk Corporation

Key material is usually AISI 1018 or AISI 1045. Table 11.4 provides the allowable stresses for these materials. Example: Select a key for the following conditions: 300 hp at 600 rpm; 3" diameter shaft, 34 " × 34 " key, 4" key engagement length. T = Torque =

300 × 63,000 hp × 63,000 = = 31,500 in-lbs rpm 600

2 × 31,500 2T = = 7,000 psi d×w×L 3 × 3/4 × 4 2 × 31,500 2T = = 14,000 psi Sc = d × h1 × L 3 × 3/8 × 4 Ss =

The AISI 1018 key can be used since it is within allowable stresses listed in Table 11.4 (allowable Ss = 7,500; allowable Sc = 15,000).

Couplings 243

Note: If shaft had been 2- 34 " diameter (4" hub), the key would be 85 " × 58 ", Ss = 9,200 psi, Sc = 18,400 psi, and a heat-treated key of AISI 1045 would have been required (allowable Ss = 15,000, allowable Sc = 30,000).

Shaft Stress Calculations Torsional stresses are developed when power is transmitted through shafts. In addition, the tooth loads of gears mounted on shafts create bending stresses. Shaft design, therefore, is based on safe limits of torsion and bending. To determine minimum shaft diameter in inches: hp × 321000 Minimum shaft diameter = 3 rpm × Allowable stress Example: hp = 300 rpm = 30 Material = 225 Brinell From Figure 11.20 at 225 Brinell, allowable torsion = 8,000 psi √ 300 × 321000 3 = 402 = 7. 38 inches Minimum shaft diameter = 3 30 × 8000 From Table 11.5, note that the cube of 7 14 " is 381, which is too small (i.e.,

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